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For full announcement see 1'st following index. 



STEAM POWER 



BY 

C. F. HIRSHFELD, M.M.E. 

FOKMEKLY PROFESSOR OF POWER ENGINEERING, SIBLEY COLLEGE, 
CORNELL UNIVERSITY 



T. C. ULBRICHT, M.E., M.M.E. 

FORMERLY INSTRUCTOR, DEPARTMENT OF POWER ENGINEERING, 

SIBLEY COLLEGE, CORNELL UNIVERSITY, MEMBER 

AMERICAN SOCIETY OF MECHANICAL ENGINEERS 



SECOND EDITION 
Revised and Enlarged 

TOTAL ISSUE FOURTEEN THOUSAND 






NEW YORK 

JOHN WILEY & SONS, Inc 

London: CHAPMAN & HALL, Limited 

1922 






Copyright, 191G, 1922, by 
C. F. HIRSHFELD and T. C. ULBRICHT 



$ y . .. , - ^ 



8/22 



DEC -SB22 

PRESS OF 

BRAUNWORTH & CO. 

BOOK MANUFACTURERS 

BROOKLYN, N. Y„ 



C1A690514 



PREFACE TO SECOND EDITION 



The comments and criticisms which have come to the 
authors during the five years since this book appeared do 
not seem to call for any radical change in arrangement or 
treatment. There does appear to be a demand for the addi- 
tion of a small amount of material. This has been met by 
the inclusion of a chapter on " Performance of Steam Power 
Equipment " in which are included treatments of those sub- 
jects which users of the book seem to feel are necessary for 
completeness. 

Other additions of minor character have been made to 
care for the development of the art since the date of the 
first edition. 

The entire text has been carefully reviewed and certain 
parts have been rewritten to make them clearer. Certain 
minor errors and misprints discovered by users have also 
been corrected. 

The authors desire to express their sincere thanks to all 
the users of this book who have assisted in this revision 
by indicating errors and omissions and points at which the 
text was not readily intelligible. 

C. F. H. 
T. C. U. 



A' 



PREFACE TO FIRST EDITION 



The following pages represent the results of an attempt 
to collect in a comparatively small book such parts of the 
field of steam power as should be familiar to engineers 
whose work does not require that they be conversant with 
the more complicated thermodynamic principles considered 
in advanced treatments. The experience of the authors 
has led them to believe that a book of this sort should give 
a correct view-point with regard to the use of heat in the 
power plant even though it does not enter deeply into the 
theoretical considerations leading up to that view-point; 
that it should supply the tools required for the solution of 
power plant problems of the common sort ; and that it should 
give sufficient description of power plant apparatus to 
make the reader fairly familiar with the more common 
types. 

Mathematical treatment of the subject has been elim- 
inated to the greatest possible extent, and anyone familiar 
with elementary algebra should be able to understand 
readily such equations as it has been deemed necessary to 
include. 

Brief explanations of physical and chemical concepts 
are given in every case in which the text required their use, 
so that those who have not studied these subjects, and those 
who have but have failed to crystallize and hold the neces- 



v i PREFACE 

sary ideas, should have little difficulty in reading the text 

understandingly. 

It is hoped that the book may prove serviceable as a 

text for steam power courses given to civil engineers in the 

various colleges and that it may also meet the needs of those 

instructing power plant operators in industrial schools. 

C. F. H. 

T. C. U. 
June, 1916. 



CONTENTS 



CHAPTER I 

PAGE 

Physical Conceptions and Units 1 

1. Matter. 2. Energy. 3. Units of Matter and Energy. 
4. Work. 5. Mechanical Energy. 6. Heat. 7. Temper- 
ature. 8. Measurement of Temperature. 9. The Unit of 
Heat Energy. 10. Specific Heat. 11. Quantity of Heat. 
12. Work and Power. 

CHAPTER II 

The Heat-power Plant 20 

13. The Simple Steam-power Plant. 14. Cycle of Events. 
15. Action of Steam in the Cylinder. 16. Hydraulic Analogy. 

CHAPTER III 

Steam 27 

17. Vapors and Gases. 18. Properties of Steam. 19. 
Generation of Steam or Water Vapor. 20. Heat of Liquid, 
q or h. 21. Latent Heat of Vaporization, r or L. 22. Total 
Heat of Dry Saturated Steam, X or H. 23. Total Heat of Wet 
Steam. 24. Heat of Superheat. 25. Total Heat of Super- 
heated Steam. 26. Specific Volume of Dry Saturated 
Steam, V or S. 27. Specific Density of Dry Saturated 

Steam, — or 8. 28. Reversal of the Phenomena Just De- 
scribed. 29. Generation of Steam in Real Steam Boiler. 
30. Gauge Pressure. 

CHAPTER IV 

The Ideal Steam Engine 43 

31. The Engine. 32. Operation of the Engine. 33. Work 
Done by the Engine. 34. Heat Quantities Involved. 35. 
Efficiency. 36. Effect of Wet Steam. 37. Application to 



viii CONTENTS 

PAGE 

a Real Engine. 38. Desirability of Other Cycles. 39. The 
Complete Expansion Cycle. 40. The Incomplete Expan- 
sion Cycle. 

CHAPTER V 

Entropy Diagram 61 

41. Definitions. 42. Temperature-Entropy Chart for 
Steam. 43. Quality from T^-Chart. 44. Volume from T</>- 
chart. 45. Heat from T0-chart. 46. The Complete T<£- 
chart for Steam. 

CHAPTER VI 

Temperature Entropy Diagrams of Steam Cycles 72 

47. Complete-expansion Cycle. 48. Area of Cycle Repre- 
sentative of Work. 49. Modifications for Wet and Super- 
heated Steam. 50. Incomplete Expansion Cycle. 51. 
Effect of Temperature Range on Efficiency. 

CHAPTER VII 

The Real Steam Engine 77 

52. Operations of Real Engine. 53. Losses in Real In- 
stallations. 54. Clearance. 55. Cushion Steam and Cylinder 
Feed. 56. Determination of Initial Condensation. 57. 
Methods of Decreasing Cylinder Condensation. 58. Classi- 
fication of Steam Engines. 59. Rotative Speeds and Piston 
Speeds. 60. The Simple D-Slide Valve Engine. 61. Engine 
Nomenclature. 62. Principal Parts of Engines. 

CHAPTER VIII 

The Indicator Diagram and Derived Values 115 

63. The Indicator. 64. Determination of I.h.p. 65. 
Conventional Diagram and Card Factors. 66. Ratio of 
Expansion. 67. Determination of Clearance Volume from 
Diagram. 68. Diagram Water Rate. 69. T^-diagram for 
a Real Engine. 70. Mechanical and Thermal Efficiencies. 

CHAPTER IX 

Compounding 141 

71. Gain by Expansion. 72. Compounding. 73. The 
Compound Engine. 74. Cylinder Ratios. 75. Indicator 
Diagrams and Mean Pressures. 76. Combined Indicator 
Diagrams. 



CONTENTS ix 

CHAPTER X 

PAGE 

The D-Slide Valve 159 

77. Description and Method of Operation. 78. Steam Lap. 
79. Lead. 80. Angle of Advance. 81. Exhaust Lap. 82. 
The Bilgram Diagram. 83. Exhaust and Compression. 84. 
Diagram for Both Cylinder Ends. 85. Piston Positions. 
86. Indicator Diagram from Bilgram Diagram. 87. Limita- 
tions of the D-slide Valve. 88. Reversing Engines. 89. Valve 
. Setting. 

CHAPTER XI 

Corliss and Other High-efficiency Engines 196 

90. The Trip-cut-off Corliss Engine. 91. Non-detaching 
Corliss Gears. 92. Poppet Valves. 93. The Una-flow En- 
gine. 94. The Locomobile Type. 

CHAPTER XII 

Regulation 216 

95. Kinds of Regulation. 96. Governor Regulation. 97. 
Methods of Varying Mean Effective Pressure. 98. Con- 
stant Speed Governing. 99. Governors. 

CHAPTER XIII 

The Steam Turbine 224 

100. The Impulse Turbine. 101. Theoretical Cycle of 
Steam Turbine. 102. Nozzle Design. 103. Action of Steam 
on Impulse Blades. 104. De Laval Impulse Turbine. 105. 
Gearing and Staging. 106. The Reaction Type. 107. Com- 
bined Types. 108. Steam Consumption of Steam Turbines. 
109. Low Pressure Turbines. 110. Steam Turbo-generators. 

CHAPTER XIV 

Condensers and Related Apparatus 261 

111. The Advantage of Condensing. 112. Measurement 
of Vacuum. 113. Conversion of Readings from Inches of 
Mercury to Pounds per Square Inch. 114. Principle of the 
Condenser. 115. Types of Condensers. 116. The Jet Con- 
denser. 117. Non-contact Condensers. 118. Vacuum 
Pumps or Air Pumps. 119. Water Required by Contact 
Condensers. 120. Weight of Water Eequired by Non-contact 



CONTENTS 



PAGE 

Condensers. 121. Relative Advantages of Contact and Sur- 
face Condensers. 122. Cooling Towers. 



CHAPTER XV 

Combustion 296 

123. Definitions. 124. Combustion of Carbon. 125. 
Combustion to CO. 126. Combustion to C0 2 . 127. Com- 
bustion of CO to CO2. 128. Conditions Determining Forma- 
tion of CO and C0 2 . 129. Flue Gases from Combustion of 
Carbon. 130. Combustion of Hydrogen. 131. Combustion 
of Hydrocarbons. 132. Combustion of Sulphur. 133. Com- 
bustion of Mixtures. 134. Composition of Flue Gases. 135. 
Temperature of Combustion, 



CHAPTER XVI 

Fuels 317 

136. Commercial Fuels. 137. Coal. 138. Coal Analyses. 
139. Calorific Value of Coals. 140. Purchase of Coals on 
Analysis. 141. Petroleum. 

CHAPTER XVII 

Steam Boilers 326 

142. Definitions and Classification. 143. Functions of 
Parts. 144. Furnaces and Combustion. 145. Hand Firing. 
146. Mechanical Grates. 147. Smoke and Its Prevention. 
148. Mechanical Stokers. 149. Rate of Combustion. 150. 
Strength and Safety of Boiler. 151. Circulation in Boilers. 
152. Types of Boilers. 153. Boiler Rating. 154. Boiler 
Efficiencies. 155. Effects of Soot and Scale. 156. Scale. 
157. Scale Prevention. 158. Superheaters. 159. Draft 
Apparatus. 



CHAPTER XVIII 

Recovery of Waste Heat 39ft 

160. Waste Heat in Steam Plant. 161. Utilization of Ex- 
haust for Heating Buildings. 162. Feed-water Heating. 



CONTENTS xi 

CHAPTER XIX 

PAGE 

Boiler-feed Pumps and Other Auxiliaries. 410 

163. Boiler-feed Pumps. 164. The Steam Injector. 165. 
Separators. 166. Steam Traps. 167. Steam Piping. 



CHAPTER XX 

Performance of Steam Power Equipment 419 

168. Meaning of Performance. 169. Determination of 
Boiler Performance. 170. Fuel Calorimeters. 171. Steam 
Calorimeters. 172. Determination of Engine Performance. 
173. Determination of Delivered Horse-power. 



STEAM POWER 



CHAPTER I 
PHYSICAL CONCEPTIONS AND UNITS 

1. Matter. The universe is generally pictured as com- 
posed of matter and energy. Matter is regarded as that 
which is possessed of mass, or as that which is possessed of 
inertia; i.e., which requires the action of force to put it in 
motion, to bring it to rest or to change its velocity. These 
definitions merely enumerate characteristics of matter; they 
do not tell what it really is. In the present state of knowledge 
it is, however, impossible to define matter in any other way. 

No experiment has yet shown that matter can be created 
or destroyed by man. It can be changed from one form to 
another, it can be given certain physical and certain chem- 
ical characteristics, more or less at will, but the actual 
quantity of matter concerned is always the same after and 
before such changes. It is customary to state this experi- 
ence in the form of a law known as the Law of the Con- 
servation of Matter, which states that the " total quantity 
of matter in the universe is constant." 

Matter is known to exist in several physical states or 
conditions of aggregation. The three most familiar are (1) 
solid, (2) liquid and (3) gaseous. In each of these states 
matter is conceived as made up of minute particles called 
molecules which in turn are apparently composed of still 
smaller parts known as atoms. These atoms can also be 
broken into parts, but for the purposes of this book it is not 
necessary to consider such further subdivision. 



2 STEAM POWER 

Experiment and mathematical reasoning seem to indi- 
cate that the molecules of all materials are in constant 
motion and that there are neutralizing attractive and repul- 
sive forces acting between them. In solids the molecules are 
apparently bound together in such a way that, although they 
are in constant motion, the external form or shape of the 
body tends to remain constant; in fact it requires the 
expenditure of force to cause a change of form. In liquids 
the molecular attraction is so altered that practically all 
rigidity disappears and the shape assumed by the liquid is 
determined by that of the surrounding surfaces, as, for 
instance, the shape of the vessel containing the liquid. In 
gases the molecules are still more free and actually tend to 
move apart as far as possible, so that a gas will spread in 
all directions until it fills any closed containing vessel. 

2. Energy. Nearly everyone has a conception of what 
is meant by the term energy, but no one yet knows what 
energy really is. It is defined as the capacity for doing work, 
or the ability to overcome resistance. A man is said to be very 
energetic or to be possessed of a great deal of energy when 
he has the ability to perform a great amount of work or 
to overcome great resistances. Matter is said to be pos- 
sessed of energy when it can perform work or overcome 
resistance. Actually, matter is not known in any form in 
which it is not possessed of energy. 

There are many different forms of energy. A body in 
motion can do work and is said to be possessed of mechani- 
cal energy. A body which we recognize as hot can do work 
at the expense of the heat associated with it and is said 
to be possessed of heat energy. Light, sound and electricity 
are all forms of energy. 

Experiment and experience have never shown that energy 
can be destroyed or created by man, but they have shown 
that one form of energy can be converted into another form 
under proper conditions. The first part of this experience 
is stated as a law known as the Law of the Conservation of 



PHYSICAL CONCEPTIONS AND UNITS 



Energy. This law states that " the total quantity of energy 
in the universe is constant" 

3. Units of Matter and of Energy. When attempts are 
made to measure the amount of anything, some unit of 
measurement is adopted. Matter is measured in numerous 
ways and many units are used. The common methods of 
measuring matter are by volume and by weight. Engineers 
in English-speaking countries use the cubic yard, the cubic 
foot or the cubic inch as units in measuring matter by vol- 
ume and they use the pound, the ounce, the grain, etc. as 
units in measuring matter by weight. 

Energy is measured in many units and, in general, there 
is a characteristic unit or set of units for each form in which 
it occurs. Thus the foot-pound is very commonly used for 
measuring mechanical energy; the British thermal unit for 
measuring heat energy; and the joule for measuring electrical 
energy. Some of these units will be defined and considered 
in greater detail in subsequent paragraphs. 

4. Work. Work is defined as the overcoming of 
a resistance through a distance. Thus, work is done 
when a weight is raised against the resistance 
offered by gravity; work is done when a spring is 
compressed against the resistance which the 
metal offers to change of shape; work is done 
when a body is moved over another against the 
resistance offered by friction. 

The unit of work is the quantity of work which 
must be done in raising a weight of one pound 
through a vertical distance of one foot. It is called 
the foot-pound. Thus, one foot-pound of work 
must be done in raising one pound one foot; two 
foot-pounds of work must be done in raising two 
pounds one foot or in raising one pound two feet. fig. i 

If a weight of one pound were suspended from 
a spring balance as shown in Fig. 1, the balance would in- 
dicate a pull or force of one pound. No work would be 




4 STEAM POWER 

done by this force as long as the weight remained stationary, 
because no resistance would be overcome through a distance. 
If, however, the same weight were slowly or rapidly raised 
a vertical distance of a foot, one foot-pound of work would 
be done. A force or pull of one pound would then have 
overcome a resistance of one pound through a distance of 
one foot. In general: 

Work in ft.-lbs. = Resistance overcome in lbs. X distance. 
= Force in Ibs.Xdistance in ft. 

so that if a force of 10 lbs. pushes or pulls anything which 
offers a resistance of 10 lbs. while that something travels 
a distance of, say, 5 ft., the work done will be given by the 
expression, 

Work =10X5, 

= 50 ft.-lbs. 

A body in falling a certain distance can do work equal to 
its weight multiplied by the distance it falls because it could 
theoretically raise an equal weight an equal distance against 
the action of gravity, and the work done upon this second 
body would be equal to its weight multiplied by the distance 
through which it was raised. 

It is very important to note that no work is done by a 
force if there is no motion; resistance must be overcome 
through a distance in order that work may be done. Thus, 
a force of 1000 lbs. might be required to hold something in 
position, that is to balance a resistance, but no work would 
be done if the body upon which the 1000-pound force acted 
did not move. Again, a weight of 50 lbs. held at a distance ol 
10 ft. above the surface of the earth would exert a downward 
push or pull equal to 50 lbs. on whatever held it in that 
position; it would, however, do no work if held in that 
position. If allowed to fall through the distance of 10 ft. 
it could do 50X10 = 500 ft.-lbs. of work. 

It is very convenient to represent graphically the action 



PHYSICAL CONCEPTIONS AND UNITS 5 

of forces overcoming resistances, that is, doing work. This 
is done by plotting points showing the magnitude of the 
force at the time that the body on which it is acting has 
traveled different distances. Thus, suppose a constant 
force of 10 lbs. pushes a body a distance of 15 ft. against a 
constant resistance of 10 lbs. The force acting on the body 
will have a value of 10 lbs. just as the body starts to move, 
a value of 10 lbs. when the body has moved 1 ft., a value 
of 10 lbs. when the body has moved 2 ft., and so on. This 
might be represented by points on squared paper as shown 






o 4 
3 















































































1 





























































































































































































































































































6 7 8 9 10 11 
Distance traveled in Feet 

Fig. 2. 



12 13 14 15 



in Fig. 2 or by a horizontal line joining those points as shown 
in the same figure. 

The work done by this force would be 10X15 = 150 
ft.-lbs. according to our previous definition. But 10X15 
is also the number of small squares under the line represent- 
ing the action of this force in Fig. 2. The number of these 
small squares then must be a measure of the work done, 
but it is also a measure of the area under the line represent- 
ing the action of the force, so that this area must be a measure 
of the work done. Each small square represents 1 lb. by 
its vertical dimension and 1 ft. by its horizontal dimension, 



6 



STEAM POWER 



so that its area must represent 1 lb. XI ft. = 1 ft. -lb. 
The total number of squares below the line equals 10 X 
15 = 150, and since the area of each one represents 1 ft.-lb. 
the total area under the line represents 150X1 = 150 
ft.-lbs. 

It is not always convenient to choose such simple scales 
as those just used. Thus it might be more convenient 
to plot the action of this force as is done in Fig. 3. Here 
the height of a square represents 2 lbs. and the width 
represents 1 ft.; the area then represents 2X1=2 ft.-lbs. 
There are 5X15 = 75 squares under the line and as each 



„ 12 
lio 

I 8 
.5 6 

o 4 


























































































































































































L 5 


> j 


5 


L 


> ( 
Dis 


tanc 


1 
3 tra 


i i 

veled 


) l 
inF 


1 
eet 


1 12 13 . 14 15 



Fig. 3. 



represents 2 ft.-lbs. the total area under the line represents 
2X75 = 150 ft.-lbs. as before. 

This is a very useful property of these diagrams and the 
area under the line representing the action of the force 
always represents the work done, no matter what the shape 
of that line. 

Thus, assume a force which compresses a spring a distance 
of 6 ins. Suppose that a force of 10 lbs. is required to com- 
press the spring 1 in., a force of 20 lbs. to compress it 2 ins., 
and so on up to a force of 60 lbs. to compress it 6 ins. 
Starting with a force of zero, the force will have to gradually 
increase as the spring is compressed, as shown by the line in 
Fig. 4. The area of each of the small squares will represent 



wx h=T2 ft - lbs - 



Under the line there is an area equal 



PHYSICAL CONCEPTIONS AND UNITS 



to 5^5 - is sma ,n squares, and the work done in compressing 

the spring must then be 18X?o = 15 ft.-lbs. 

5. Mechanical Energy. Any body which exists in such 
a position or location that it could do work by dropping or 
falling is said to be possessed of potential mechanical energy, 
or of mechanical energy due to position. As long as it 
remains in this position, it cannot do work at the expense 
of this energy, but, if allowed to fall, it could do so. The 




Fig. 4. — Graph Showing Action of Spring. 

work it could do would be equal to the product of its weight 
by the distance it could fall and the potential energy it 
possesses before starting to fall is measured by this work. 
Thus, a body weighing 40 lbs. located 10 ft. above the surface 
of the earth could do 40X10 = 400 ft.-lbs. of work in falling, 
and, therefore, it is said to be possessed of 400 ft.-lbs. of 
potential energy before it starts to fall. 

If in falling it raises a weight equal to its own (theo- 
retically) through a distance equal to that through which 
it falls (theoretically), it will have used up 400 ft.-lbs. of 
energy in doing 400 ft.-lbs. of work upon the body raised 



8 STEAM POWER 

and will no longer be possessed of that amount of potential 
energy. The body which has been raised will, however, 
have an equal amount of energy stored in it and will in turn 
be able to do 400 ft. -lbs. of work if allowed to fall a distance 
of 10 ft. 

If the body assumed above falls through a distance of 
10 ft. without raising another body or doing an equivalent 
amount of work in some other way, it acquires a high 
velocity. When it arrives at the bottom of the fall of 10 
ft., it certainly does not possess the 400 ft. -lbs. of potential 
energy which it had before dropping nor has it done work 
at the expense of that energy. Moreover, the energy could 
not have been destroyed because it is indestructible. The 
only conclusion is that it must still be possessed of this 
energy in some way. At the end of the fall it has lost its 
advantageous position, but it has acquired a high velocity, 
and experience shows that if brought to rest it can do 
work upon that which brings it to rest equal to what 
it could have done in raising a weight as previously 
described. 

At. the end of its fall and before being brought to rest, 
the body is therefore said to be possessed of energy by virtue 
of its velocity, and this form of energy is called kinetic 
mechanical energy. The kinetic energy will be exactly equal 
to the 'potential energy which disappeared as the body fell. 

Any body which is moving is possessed of kinetic energy 
because it can do work on anything which brings it to 
rest. This energy is expressed by the equation, 

1 W 

Kinetic Energy in ft.-lbs.=-Xw^X V 2 , 

in which 

W = the weight of the moving body in pounds. 
F = the velocity in ft. per second, and 
32.2 = a gravitational constant commonly represented by g. 



PHYSICAL CONCEPTIONS AND UNITS 9 

6. Heat. One of the most familiar forms of energy is 
heat, which manifests itself to man through the sense of 
touch. In reality every body with which man is familiar 
possesses an unknown amount of heat energy and it is 
assumed that this heat energy is in some way associated 
with the motions and relative positions of the molecules and 
their constituents. 

For this reason heat is often described as molecular 
activity and is regarded as energy stored up in a substance by 
virtue of its molecular condition. Heat energy can be made 
to perform work in ways which will be discussed later and 
this is proof that it is a form of energy and not a material 
substance, as was once supposed. 

Heat is observed and recorded by its effects on matter, 
producing changes in the dimensions or volumes of objects; 
changes of internal stress; changes of state, as ice to water 
and water to steam; changes of temperature; and electrical 
and chemical effects. 

Neglecting certain atomic phenomena not yet well under- 
stood, the probable source of all heat energy appearing on 
the earth is the sun. Heat, however, may be obtained 
from mechanical and electrical energy; from chemical 
changes; from changes of physical state; from the internal 
heat of the earth. 

7. Temperature. Man early realized that under certain 
conditions bodies felt " hotter " than under other conditions 
and gradually came to speak of the " degree of hotness " as 
the temperature of the body. It was later realized that what 
was really measured as the " hotness " or intensity of heat 
or temperature of a body was the ability of that body to trans- 
mit heat to others and that it had no connection with quantity 
of heat. 

Thus, if the temperature of two adjacent bodies happened 
to be the same, one of them could not lose heat by trans- 
mitting it to the other, but if the temperature of one 
happened to be higher than that of another, the body at 



10 STEAM POWER 

higher temperature would always lose heat to the one at 
lower temperature. 

As a means of measuring temperature certain arbitrary 
scales have been chosen. The centigrade scale of tempera- 
ture, for instance, is based upon the temperatures of melting 
ice and boiling water under atmospheric pressure. The tem- 
perature difference between boiling water at atmospheric 
pressure and melting ice at atmospheric pressure is arbi- 
trarily called one hundred degrees of temperature, and the 
temperature of the melting ice is called zero, making that 
of the boiling water 100 degrees. 

Any body which has such a temperature that it will not 
give heat to, or take heat from, melting ice is said to be at 
a temperature of zero degrees centigrade, represented as 0° C. 
Similarly, any body in such a condition that it will not give 
heat to or take heat from water boiling under atmospheric 
pressure is said to have a temperature of 100° C. A body 
with a temperature exactly half way between these two 
limits would then be said to have a temperature of 50° C. 

8. Measurement of Temperature. The temperatures of 
bodies could be determined by bringing them in contact with 
such things as melting ice and boiling water and determining 
whether or not a transfer of heat occurred, but this would 
be a very cumbersome and unsatisfactory method. As a 
consequence many other means have been devised for the 
measurement of temperature. 

One of the most common and convenient methods in- 
volves the use of what are known as mercury thermometers. 
These depend upon the fact that the expansion of mercury 
with changing temperature is very uniform over a wide 
temperature range. Thus, if mercury expands a certain 
amount when its temperature is raised from that of melting 
ice to that of boiling water, i.e., 100° C, it will expand just 
half as much when its temperature is raised half as high, 
and one-quarter as much when its temperature is raised one- 
quarter of the range from 0° to 100° C. 



PHYSICAL CONCEPTIONS AND UNITS 



11 



n 



The thermometer is made by enclosing a small quan- 
tity of mercury in a glass tube fitted with a bulb at 
one end, as shown in Fig. 5. The lower end 
of the thermometer is immersed in melting 
ice and the point on the stem which is 
reached by the top of the mercury column is 
marked and labelled 0° C. The thermometer 
is then immersed in the steam from water boil- 
ing under atmospheric pressure and the point 
reached by the top of the mercury column is 
marked and labelled 100° C. The distance be- 
tween the two marks is then divided into one 
hundred parts and each represents the distance 
which the end of the column of mercury will 
move when its temperature changes one centi- 
grade degree. 

It is customary to extend this same scale 
below 0° and above 100°, carrying it, on ex- 
pensive thermometers, as far in each direction 
as the approximation to a constant expansion 
on the part of the mercury and to constant 
properties of the glass justifies. 

The temperature of a body can then 
be found by placing the thermometer in 
or in contact with that body and noting 
the point reached by the end of the 
mercury column. The division reached 
gives the temperature directly. 

The centigrade scale just described is 
the one commonly used by scientists the 
world over, but engineers in this country 
more often use what is known as the 
Fahrenheit scale, This is so chosen that 
the temperature of melting ice is called 
32° F. and the temperature of water boil- 
ing under atmospheric pressure is called 212° F. There are 



i 



Fig. 5.— Mer- 
cury Ther- 
mometer. 



c.° 



i 



o 

-1778 c 



« 



Fig. 6. — Comparison 
of Centigrade and 
Fahrenheit Scales, 



12 



STEAM POWER 



thus 180° on this scale for the same temperature difference 
as is represented by 100° on the centigrade scale. The 
relation between the two scales is shown diagrammatically 
in Fig. 6. It is apparent that the temperature of a body 
at 0° C. will be 32° F. and that of a body at 0° F. will be 
-17.8° C. 

Since 100 centigrade degrees are equal to 180 Fahren- 
heit degrees, it follows that 



P 180 9 
C = 100 = 5 K 



and that 



1 F -T80~9 °- 



(1) 



(2) 



Therefore, if t F and t c represent temperatures on the Fahren- 
heit and centigrade scales respectively, 



tF — -=tc ~T~o2i . 



and 



9 



(fc-32) 



(3) 



(4) 



Cent . ^ Abs. — ^Fahr. 



100 






073 



273" 491.4 

459.4' 



There is still another temperature scale of great impor- 
tance. It is known as the absolute 
scale and temperatures measured 
on it are spoken of as absolute tem- 
peratures. The zero on this scale 
is located at -273° C. or 273 centi- 
grade degrees below centigrade 
zero, or, what is the same thing, 
at -459.4° F., or 459.4 Fahrenheit 
degrees below Fahrenheit zero. The 
degrees used are either centigrade 
or Fahrenheit, as convenient, so 
that there are absolute tempera- 
tures expressed in centigrade de- 
grees above absolute zero and there are absolute tempera- 



-273 

Fig. 7. — Comparison of Ab- 
solute and Ordinary 
Temperature Scales. 



PHYSICAL CONCEPTIONS AND UNITS 13 

tures expressed in Fahrenheit degrees above absolute zero. 
The relations between the various scales are shown dia- 
grammatically in Fig. 7. 

It is apparent from this diagram that, 



and that 



7V = £f+460 (approximately) . . . (5) 
T c = t c +27S (6) 



if TV and T c represent absolute temperatures and if the 
number 459.4 is rounded out to 460, as is commonly 
done. 

9. The Unit of Heat Energy. The unit used in the 
measurement of heat energy in the United States is the 
British Thermal Unit (abbreviated B.t.u). It is defined as 
the quantity of heat required to raise the temperature of one 
pound of pure water one degree Fahrenheit. In order to 
make the definition very exact it is necessary to state the 
temperature of the water before the temperature rise occurs, 
because it requires different amounts of heat to raise the 
temperature of a pound of water one degree from differ- 
ent initial temperatures. For ordinary engineering pur- 
poses, however, such refinements generally may be omitted. 

Many experimenters have shown that heat energy and 
mechanical energy are mutually convertible, that is, the one 
can be changed into the other. When such a change occurs 
no energy can be lost since energy is indestructible, and it 
follows that, if one form is changed into the other, there 
must be just as much energy present after the change as 
there was before. 

As the units used in measuring the two forms of energy 
are very different and as it is often necessary to express 
quantities of energy taking part in such conversions, it is 
desirable to determine the relations between these units. 
This was first accurately done by Joule, who showed that one 
British thermal unit of heat energy resulted from the con- 



14 STEAM POWER ' 

version of 772 ft.-lbs. of mechanical- energy. Later experi- 
menters have shown that the number 778 more nearly 
expresses the truth than does the number 772 and the larger 
value is now known as Joule's Equivalent. 

Expressed mathematically, the relation between the units 
is 

1 B.t.u.= 778 ft.-lbs (7) 

lft.-lb.=^B.t.u (8) 

10. Specific Heat. The specific heat of a substance is 
defined as that quantity of heat which is used up or recovered 
when the temperature of one pound of the material in question 
is raised or lowered one degree. Its numerical value depends 
upon the specific heat of water since the quantity of heat is 
measured in units dependent upon the amount required to 
raise the temperature of water. The specific heat of water 
is, however, very variable, as shown by the values given in 
Table L, and it is therefore evident that exact numerical 
values of specific heats can only be given when the definition 
of the B.t.u. is exactly expressed. 

The specific heats of all real substances vary with tem- 
perature and the values commonly used are either rough 
averages or are those determined by experiments at one 
temperature. For most engineering purposes errors arising 
from this source may, however, be neglected. 

From the definition of specific heat it follows that : 



C ~W{t2-hY (9) 



in which 



C = a mean or average specific heat over a range of tem- 
perature from t\ to t2, and 

Q = the heat supplied to raise the temperature of W 
pounds of material from t\ to fo. 



PHYSICAL CONCEPTIONS AND UNITS 



15 



TABLE I 
Specific Heats of Water. 

(Value at 55° F. taken as unity) 



Temp. F°. 


Spec. Ht. 


Temp. F°. 


Spec. Ht. 


20 


1.0168 


250 


1.045 




30 


" 1.0098 


400 


1.064 




40 


1.0045 


450 


1.086 




50 


1.0012 


500 


1.112 




60 


0.9990 


510 


1.117 




70 . 


0.9977 


520 


1.123 




80 


0.9970 


530 


1.128 




90 


0.9967 


540 


1.134 




100 


0.9967 


550 


1.140 




120 


0.9974 


560 


1.146 




140 


0.9986 


570 


1.152 




160 


1.0002 


580 


1.158 




180 


1.0019 


590 


1.165 




200 


1.0039 


600 


1.172 




220 


1.007 








240 


1.012 








260 


1.018 








280 


1.023 








300 


1.029 









* Values taken from Marks and Davis, " Steam Tables and Diagrams," p. 68. 

ILLUSTRATIVE PROBLEMS 

1. Given: Sp. ht. of iron=0.113, of aluminum =0.211; Initial 
temp. =150° F. Temp, range (k-U) =100° F. 

If 1 lb. of iron and 1 lb. of aluminum are cooled through this 
temperature range, how much more heat is lost in one case than 
in the other? 

Q a i = IFCai(*2-* 1 )==lX.21lX100=21.1 B.t.u. 

Q lT = WCir(t 2 -t 1 )=lX. 113X100 = 11.3 B.t.u. 

Difference 9.8 B.t.u. 

2. If the difference obtained in Prob. 1 were used to heat up 
5 lbs. of silver, with a specific heat equal to 0.057, what would be 
the temperature range through which it would be raised? 

Q =9.8 =5X0.057(^2 -^) =0.285(£ 2 -h) 
/. k-*i=34.4° F. 



16 STEAM POWER 

3. If the initial temperature of the silver in Prob. 2 were 150° F e 
what would be the final absolute temperature Fahr.? 

k =fc+34.4° = 150+34.4 = 184° (approximately). 

T 2 =460+184 =644° F. Abs. 

4. 100 lbs. of water in a 20-lb. tank of iron, both at 60° F., 
are placed in salt brine at 0° F. The water becomes ice at' 32° F. 
and the temperature of the ice is lowered to 26° F., the brine being 
raised to 26° F. Sp. ht. water = 1.0; Sp. ht. ice =0.5; Sp. ht. 
iron =0.113; Sp. ht. brine =0.8; and 143 B.t.u. per pound of water 
must be removed to convert liquid water at 32° F. to ice at the 
same temperature. What weight of brine is required? 

100[l(60-32) + 143+.5(32 -26)]+20X0.113(60 -26) 
= 17X0.8(26-0) 
W = 840 lbs. of brine. 

11. Quantity of Heat. It is impossible to determine the 
total quantity of heat in or " associated with " a substance, 
because no means of removing and measuring all the heat 
contained in any real material have ever been devised. 
Since, however, the engineer is concerned with changes of 
heat content ratber than with the total amount of heat 
contained, this fact causes him no difficulty. 

For convenience in figuring changes of heat content, 
it is customary to assume some arbitrary starting 'point or 
datum and to call the heat in the material in question zero 
at that point. 

Thus, for example, if it were necessary to figure heat 
changes experienced by a piece of iron weighing 5 lbs. and 
having a specific heat of 0.1138, and the temperature of this 
iron never dropped below 40° F. under the conditions exist- 
ing, this temperature might be taken as an arbitrary starting 
point above which to figure heat contents. If the iron were 
later found at a temperature of 75° F., " the heat content 
above 40° F." would be said to be 

Q = CW (t 2 -h) =0.1138X5(75-40) =2.27 B.t.u. 



PHYSICAL CONCEPTIONS AND UNITS 17 

This type of formula can only be used when the sub- 
stance does not change its state between the limits of tem- 
perature concerned. In the case of water which might 
change to steam during such a rise of temperature, it might 
be necessary to include other heat quantities in the cal- 
culations, as shown in a later chapter. 

12. Work and Power. Since steam engines are designed 
for the purpose of converting the heat energy contained in 
fuel into mechanical energy which may be used to perform 
work, it will be necessary to consider the units used in 
measuring work and power. 

Work was defined in a previous paragraph as the over- 
coining of a resistance through a distance, by the application 
of a force; that is, a force expressed in pounds, multiplied by 
the distance in feet through which the force acts, gives a product 
expressed in foot-pounds. 

The amount of work performed in a unit of time is termed 

power, which may be defined as the rate of doing work, 

Therefore, 

.p. Force X Distance ,.,_,. 

Power = —. j—. -. . . . (10) 

lime (mm. or sec.) 

The unit of power used by steam engineers is the horse- 
power, which is equivalent to the performance of 33,000 
ft.-lbs. of work per minute, or 550 ft. -lbs. of work per second, 
or 1,980,000 ft.-lbs. per hour. Therefore, the horse-power 
developed by any mechanism is 

, ft.-lbs. of work per min. ,.,.,. 

h - p - = 3^000 • • • • (U) 

Since 33,000 ft.-lbs. of work can be accomplished only by 

the expenditure of 33,000 ft.-lbs. of energy and since one 

B.t.u. of energy is equal to 778 ft.-lbs., it follows that 33,000 

33 000 
ft.-lbs. of work must be the equivalent of ' =42.41 B.t.u. 

77o 

It is customary to speak of power in terms of horse- 



18 STEAM POWER 

power-hours. One horse-power-hour means the doing of 
work equivalent to one horse-power for the period of one 
hour, or the doing of work at the rate of 33,000 ft.-lbs. per 
minute for an hour. A horse-power-hour is therefore equiva- 
lent to 33,000X60 = 1,980,000 ft.-lbs. As 33,000 ft.-lbs. 
are equivalent to 42.41 B.t.u., it follows that 42.41X60 = 
2544.6 or about 2545 B.t.u. are the equivalent of one horse- 
power-hour. 

The number 2545 should be memorized as it is very often 
used in steam-power calculations. If an engine could deliver 
one horse-power-hour for every 2545 B.t.u. it received, it 
would be working without losses of any kind; that is, all 
the heat energy entering it would leave it in the form of 
useful mechanical energy. It will be shown later that this 
is impossible even in the most perfect or ideal engine. 

REVIEW PROBLEMS 

1. Express 32° F. in degrees centigrade. 

2. Express 150° F. in degrees centigrade. 

3. Express 250° C. in degrees Fahrenheit. 

4. Express the results of problems 1, 2 and 3 in absolute values. 

5. What is the heat equivalent of 233,400 ft.-lbs. of work? 

6. Find the heat supplied 10 lbs. of water when its temperature 
is raised from 20° F. to 160° F., assuming the mean specific heat 
over this range to be 0.997. 

7. Find the temperature change of 2 lbs. of lead (sp. ht. 0.0314) 
when 20 B.t.u. are added. 

8. How many B.t.u. must be abstracted to lower the tem- 
perature of 15 lbs. of water from 212° F. to 32° F., assuming the 
specific heat of water to be unity? 

9. Find the weight of water which will have its temperature 
tripled in value by the addition of 250 B.t.u., the final temperature 
being 150° F. Assume specific heat unity. 

10. The specific heat of a piece of wrought iron is 0.113 and of 
a given weight of water is 1.015. 1 cu. ft. of water weighs approxi- 
mately 62.5 lbs. Find the increase in temperature of 4 cu. ft. of 
water when a common temperature of 65° F. results from placing 
in the water a piece of iron weighing 15 lbs. at a temperature of 
900° F. 



PHYSICAL CONCEPTIONS AND UNITS 19 

11. Find the final temperature of the mixture, when 100 lbs. 
of iron (sp. ht. =0.113), at a temperature of 1200° F. are immersed 
in 300 lbs. of water (sp. ht. 1.001) at a temperature of 50° F. 

12. Five pounds of silver (sp. ht. =0.057) at 800° F. are im- 
mersed in water at 60° F., resulting in a final temperature of 85° F. 
Assume Sp. ht. water = 1. What weight of water is necessary? 

13. An engine is developing 10 horse-power. Express this in 
ft .-lbs. of work done per minute and find the amount of heat 
energy equivalent to this quantity of mechanical energy. 

14. A pump raises 1000 lbs. of water 50 ft. every minute. 
How much work is done? Find the equivalent horse-power. 

15. An engine develops 1,980,000 ft. -lbs. of work at the fly- 
wheel per minute. 

(a) Find the horse-power developed. 

(b) If this engine operated in this way for an hour, how many 
horse-power hours would it make available? 

(c) What would be the equivalent of this number of horse- 
power hours in British thermal units? 



CHAPTER II 
THE HEAT-POWER PLANT 

13. The Simple Steam-Power Plant. The various pieces 
of apparatus necessary for the proper conversion of heat 
energy into mechanical power constitute what may be 
called a " Heat-Power Plant," just as the apparatus used 
in obtaining mechanical energy from moving water is called 
an hydraulic or water-power plant. Heat-power plants are 
distinguished as " Steam-Power Plants"; " Gas-Power 
Plants"; etc., according to the way in which the heat of 
the fuel happens to be utilized. 

The apparatus around which the plant as a whole centers, 
that is, the apparatus in which heat energy is received and 
from which mechanical energy is delivered, is termed the 
engine or prime-mover. This heat engine may use steam 
generated in boilers and may require certain apparatus, such 
as condensers, pumps, etc., for proper operation; or it may 
use gas, generated in gas-producers requiring coolers, 
scrubbers, tar extractors and holders, depending upon the 
class of fuel used and upon certain commercial considera- 
tions. Again, the power-plant may simply contain an in- 
ternal-combustion engine using natural gas, gasoline or oil, 
a type of plant which is now very common. 

But whatever type of plant is used, a general method of 
operation is common to all. Heat energy in fuel is constantly 
fed in at one end of the system and mechanical energy is 
delivered at the other end. The steam-power plant will be 
briefly described in the following paragraphs, showing the 
cycle of events with the attendant losses through the 
system. 

20 



THE HEAT-POWER PLANT 



21 




22 STEAM POWEE 

In Fig. 8 is shown a simple steam-power plant which con- 
verts into mechanical energy part of the heat energy, origi- 
nally stored in coal, by means of a prime-mover called a 
steam-engine. The main pieces of apparatus used in this 
type of plant are the steam-boiler; the steam-engine; the 
condenser; the vacuum pump; and the feed-pump. The 
energy stream shows the various losses occurring through- 
out the plant. These losses cause the " delivered energy " 
stream to be only a small fraction of the total heat sent into 
the system. 

14. Cycle of Events. 1. Fuel is charged on the grate 
under the boiler, where it is burned with the liberation of a 
large amount of energy. Air is drawn or forced through 
the grates in proper proportions to support this combustion. 
The hot gases resulting pass over the tubes, in a definite 
path set by the baffle plates, so that the largest possible 
amount of heating surface may be presented to the products 
of combustion. 

There are certain losses accompanying this operation, 
such as radiation, loss of volatile fuel passing off unburned, 
loss of fuel through the grate, and loss of heat through the 
excess air which must always be supplied to insure com- 
bustion. 

2. That part of the heat in the gases which is not lost 
by radiation from the boiler and in the hot gases flowing up 
the stack passes through the heating surfaces of the boiler 
to the water within. From 50 to 80 per cent of the 
total heat energy in the fuel passes through the heating 
surfaces and serves to raise the temperature of the water 
to the boiling point at the pressure maintained, and to con- 
vert this water into steam according to the requirements. 

3. Having obtained steam within the boiler, it is led 
through a system of pipes to a steam engine, where some of 
the heat stored in the steam is converted into mechanical 
energy by the action of that steam against a piston. The 
steam is then discharged, or exhausted, from the engine 



THE HEAT-POWER PLANT 23 

at a much lower temperature and pressure than when it 
entered. 

From 5 to 22 per cent of the available heat in the 
steam is converted into mechanical energy in the engine 
cylinder, and because of frictional and other losses occurring 
in the mechanism, only from 85 to 95 per cent of this 
energy is turned into useful work at the fly-wheel. 

4. In some plants, known as non-condensing plants, 
the exhaust steam, which still contains the greater part of 
all the heat received in the boiler, is discharged to the atmos- 
phere and represents a complete loss. In others, known as 
condensing plants, the exhaust steam is led to a condenser, 
where^ it is condensed by cold water, which absorbs and 
carries away the greater quantity of the heat not utilized 
in the engine. The • condensed steam or " condensate " is 
then either discharged to the sewer or transferred by means 
of a vacuum-pump to the hot-well, from which it is drawn 
by means of the feed-water-pump, raised to the original 
pressure of the steam, and returned to the boiler. Here 
it is again turned into steam and the cycle of operations 
outlined above is repeated. Naturally there is some loss 
due to evaporation and leaks throughout the system, so 
that " make-up " water must be supplied. 

The series of events just described constitutes a complete, 
closed cycle of operations, wherein the water is heated, 
vaporized, condensed and returned to the boiler, having 
served only as a medium for the transfer of heat energy 
from fuel to engine and the conversion of part of that 
energy within the cylinder. The water in such a case is 
known as the working substance. 

It is often more convenient to discard the working sub- 
stance after it leaves the cylinder, as suggested above in 
the case of a non-condensing plant; or, as in the case of 
a gas engine, where a new supply of working substance 
must be supplied for each cycle, because the burned gases 
of the previous cycle cannot be used again. 



24 STEAM POWER 

15. Action of Steam in the Cylinder. In order to pre- 
pare for the more detailed discussion of the action of the 
steam in the engine cylinder, to be taken up in a later 
chapter, a brief outline of the events occurring within the 
prime-mover will be considered at this point. 

Steam enters the cylinder through some kind of an 
admission valve, and acts upon the piston, just as the latter 
has approximately reached one end of its stroke and is 
ready to return. The heat-energy stored up in the steam 
causes it to expand behind the piston, thereby driving the 
latter out and performing work at the fly-wheel. At about 
90 or 95 per cent of the stroke, the exhaust valve opens, 
and the steam begins to exhaust, the pressure within the 
cylinder dropping almost to atmospheric or to that main- 
tamed in the condenser by the time the piston has 
reached the end of its stroke. . On the next or return stroke 
the remaining steam is forced out through the exhaust 
port, until, at some point before the end of the piston 
travel, the exhaust valve closes, and the low-pressure steam 
trapped in the cylinder is compressed into the clearance 
space so that its pressure rises. Admission then occurs, and 
the cycle is repeated. 

The diagram given in Fig. 9 illustrates the operation of 

steam within the cylinder. 

*^\ ear all c<3 ^ ^ _ ^ rpj^g diagram jg plotted 

between pressures of steam 
within the cylinder as 
ordinates and correspond- 



~~*~ c . ' Cut-off (Closing of Admission Valve) 





Release (Opening of 
Exhaust Valve) 

%- Ex n f^5^. in S P iston positions as 



■fj~ ^Atmospheric Pressure^ - absdSSaS 



7 piston positions Th e method of obtain- 

Closing of Exhaust Valve 

™ A a- -r, • t j- i ing such a diagram, known 

Fig. 9. — -Steam Engine Indicator & ° ' 

Diagram. as an indicator-diagram, 

will be fully described in a 

later chapter. Since vertical ordinates represent pressure 

in pounds per square inch, and horizontal abscissas renre- 



THE HEAT-POWER PLANT 



25 



sent feet moved through by the piston, the product of 
these two must be work. But the product of vertical by 
horizontal distances must also give area. Therefore, by 



Source of Water 
at High Head 




Energy Supplied 



Water 

Useful Energy- 
Motor f f Made Available 



<- Energy Discharged 



eceiver of Discharged 
Water at Low Head 



(a) 



Source of Heat 
at High Temp. 



IVvP 



& -Energy Supplied" 



High|Temp. 
tjabovej datum 




Heat 
Engine 



Useful Energy 
Made Available 



Low|Temp. 
1 2 above datum 



-Energy Discharged 

[Receiver of Discharged 
Heat at Low 
Temperature 



(6) 
Fig. 10. — Hydraulic Analogy. 



finding the area enclosed within the bounding lines of the 
cycle and multiplying this by a proper factor, the foot- 
pounds of work developed within the cylinder can be 
determined. 

16. Hydraulic Analogy. The operation of heat-engines 
is analogous to that of water-wheels. A water-wheel de- 



26 STEAM POWER 

velops mechanical energy by receiving water under a high 
head, absorbing some of its energy, and then rejecting the 
fluid under a low head. Similarly, the heat-engine receives 
heat energy at a high temperature (head), absorbs some of 
it by conversion into mechanical energy, and then rejects 
the rest at a low temperature (head) . 

The analogy can be carried still further. The water- 
wheel cannot remove all the energy from the water, nor 
can the heat-engine remove all the heat-energy from the 
working substance. There is a certain loss in the material 
discharged in both cases and this cannot be avoided. 

This analogy is illustrated diagrammatically in Fig. 10 
(a) and (6) in which the widths of the streams represent 
quantity of energy. 



CHAPTER III 
STEAM 

17. Vapors and Gases. When a solid is heated, under 
the proper pressure conditions, it ultimately melts or fuses 
and becomes a liquid. The temperature at which this 
occurs depends upon the particular material in question 
and upon the pressure under which it exists. Ice, which 
is merely solid water, melts at 32° F. under atmospheric 
pressure, while iron melts at about 2000° F. under atmos- 
pheric pressure. 

When a liquid is heated, it ultimately becomes a gas, 
similar to the air and other familiar gases. If this gas is 
heated to a very high temperature and if the pressure under 
which it is held is not too great, it very nearly obeys certain 
laws which are simple and which are called the laws of ideal 
gases. 

When the material is in a state between that of a liquid 
and that in which it very nearly obeys the laws of ideal 
gases, it is generally spoken of as a vapor. This term is 
used in several different ways and with several different 
modifying adjectives which will be explained in greater 
detail in later sections. 

18. Properties of Steam. Of the many vapors used by 
the engineer, steam or water vapor is probably the most 
important, because of the ease with which it can be formed 
and also because of the tremendous field in which it can 
be used. It is generated in a vessel known as a steam boiler y 
which is constructed of metal in such a way that it can 
contain water, and that heat energy, liberated from burning 
fuel, can be passed into the water, converting part or all of 
ft. into water vapor, that is, into steam. 

27 



28 



STEAM POWER 



The properties of water vapor must be thoroughly under- 
stood before the steam engine and steam boiler can be 
studied profitably. Probably the easiest way of becoming 
familiar with these properties is to study the use made of 
heat in the generation of steam from cold water. 

19. Generation of Steam or Water Vapor. For the pur- 
poses of development, assume a vessel of cylindrical form, 
fitted with a frictionless piston of known weight, as shown 
in Fig. 11, (a) and (b), the whole apparatus being placed 
under a bell-jar in which a perfect vacuum is maintained. 



IIH I hlli l HjB 



' — ~jf 



(a) 



m 



!&%£■ 



(b) 



— ] . .:..;,/ 



To Vacuum 
Pump 



Fig. 11. — Formation of Steam at Constant Pressure. 



Assume one pound of water in the cylinder, with the piston 
resting on the surface of the liquid. There will be some 
definite pressure exerted by the piston upon the surface of 
the liquid, and its value will be determined entirely by the 
weight of the piston. 

It is convenient in engineering practice to refer all 
vaporization phenomena to some datum temperature, and 
since the melting point of ice, 32° F., is a convenient refer- 
ence point, it is used as a standard datum temperature, in 
practically all steam-engineering work. Therefore, assuming 
the water in the jar to be at 32° F., if heat is applied the 
temperature of the liquid will rise approximately 1° F. 



STEAM 



29 



for every B.t.u. of heat added, since the specific heat of 
water is approximately unity. 

Experiment shows that for each pressure under which the 
water may exist some definite temperature will be attained at 
which further rise of temperature will cease and the liquid will 



ozu 
















































/ 






















































/ 






480 
















































/ 




















































/ 








440 














































/ 






















































/ 








a 












































j 


/ 








o 












































/ 










5 3b0 

3 












































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u 6 ^ 










































/ 




















































/ 












a 










































/ 












3 

o 








































/ 














Pi 4iU 

1 








































/ 














<D 








































/ 














at 






































/ 
















£J 160 






































/ 




















































/ 






















































/ 




















































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y 



















































































40 80 120 160 200 240 280 320 3(30 400 440 480 52 
Temperature -F.° 

Fig. 12. — Pressure-Temperature Relations, Saturated Water Vapor. 



begin to change to a vapor, that is, to vaporize. The tem- 
peratures at which vaporization occurs at different pressures 
are called the temperatures of vaporization at those pressures. 
The temperatures of vaporization of water are plotted 
against pressure in Fig. 12. It should be noted that the 
values of vaporization temperature increase very rapidly 
for small pressure changes in the case of low pressures, but 



30 STEAM POWER 

that, for the higher pressures, the variation of temperature 
is very small for enormous variations of pressure. This 
fact is of great importance in steam engineering. 

The temperatures of vaporization are tabulated with 
other properties of water vapor in so-called steam tables and 
are constantly referred to by engineers. An example of 
such a table is given on pp. 392 to 399. 

Returning now to the apparatus under discussion, as 
heat is supplied, the temperature of the water will rise from 
32° F. until it reaches the temperature of vaporization cor- 
responding to the pressure exerted upon the water by the 
piston. When this temperature is reached vaporization will 
begin, and if sufficient heat is supplied, will continue without 
change of temperature until the water is entirely converted 
into vapor. 

Up to the time at which vaporization starts the volume 
of the water will change very little, so that the piston will 
be raised only a negligibly small amount and practically no 
work will be done upon it by the water. On the other hand, 
when vaporization occurs the volume of the material will 
change by a very large amount and the piston will be 
driven out (raised) against the action of gravity. That is, 
work will be done by the steam in driving the piston out 
during the increase in volume which accompanies vaporiza- 
tion. 

It is found that a very great quantity of heat is used up 
during the process of vaporization despite the fact that no 
temperature change occurs. This is described by saying that 
the heat which is supplied during this period becomes latent, 
that is, not apparent, and the quantity of heat is therefore 
spoken of as the latent heat of vaporization. It is assumed 
to consist of two parts, that used for separating the liquid 
molecules against their attractive forces and that used for 
doing the work which is done upon the piston as it is 
moved upward. The former is called the internal latent 
heat because it is used for doing internal or intermolecular 



STEAM 31 

Work; the latter is called external latent heat because it is 
used for the doing of external work. 

It is to be noted that the internal latent heat may be 
assumed to be tied up in some way within the molecular 
structure of the material and hence to be in the steam. 
The external latent heat, on the other hand, is used up as 
fast as supplied for the purpose of driving the piston out 
against the action of gravity. When the piston has been 
raised to any point, the energy used in raising it is not in 
the steam, but is stored as potential energy in the piston. 
To get it back the piston must be allowed to drop. The 
term " external " is therefore well chosen; the external 
latent heat is in no sense in the steam; it is stored in 
external bodies or mechanism. 

After the constant temperature vaporization is complete, 
the further addition of heat will again cause a rise of tem- 
perature and a gradual increase of volume. Such raising 
of the temperature of steam already formed is called super- 
heating and results in carrying the vapor nearer and nearer 
to the condition in which it very nearly obeys the laws of 
ideal gases. Since an increase of volume accompanies super- 
heating, the molecules of the vapor must move farther and 
farther apart as superheating progresses. 

Vapor in the condition in which it is formed from the 
liquid and which has the same temperature as the liquid 
from which it was formed is called saturated vapor. This 
term can be pictured as meaning that the maximum number 
of molecules of vapor are packed into a given space; the 
addition of heat to saturated vapor would cause superheating 
and the separation of the molecules so that fewer could be 
contained in a given space. 

20. Heat of Liquid, q or h. Returning once more to 
the start of the process described in the preceding section, 
heat was added to water initially at 32° F. until the tem- 
perature of vaporization corresponding to the existing pres- 
sure was attained. The heat added during this period is 



32 STEAM POWER 

called the heat of the liquid, and is usually designated by the 
letters q or h. If the mean specific heat of water at con- 
stant pressure (C p ) for the temperature range under consid- 
eration were constant, q would be given by the equation 

q = C p (t v -32) 

in which U is the temperature of vaporization, and if C p 
were equal to 1 , it would follow that 

q = t v -S2. (12) 

Therefore, if water boils under a pressure of 50 lbs. at 
a temperature, read from the steam tables, of 281° F., it 
would follow that for 50 lbs. 

g = 281-32 = 249 B.t.u. 

But the steam tables (see p. 394) for this pressure (50 
lbs.) give q = 250.1 B.t.u., indicating, as was shown in 
Chap. I., that the specific heat of water does not remain 
constant, and for this case the mean value must have been 
approximately 1.004 as indicated by the following calcula- 
tion. 

q = C p 0,-32) or 250.1 = C P X 249 
so that 

C '=W = L004+ 

Hence it is always advisable to use the steam table 
values of q, except for very approximate calculations. 

21. Latent Heat of Vaporization, r or L. The heat 
supplied during the period of vaporization has already been 
referred to as the latent heat of vaporization and has been 
divided into internal and external latent heats. 

The internal latent heat is generally designated by p or 
by I and the external latent heat by the group of letters 
APu or by E. The group A Pu merely represents the prod- 



STEAM 33 

uct of pressure, P, by volume change during vaporization, 
u, and by the fraction -^ which is represented by A. The 
product of the first two terms gives external work in foot- 
pounds during vaporization, and dividing this by 778 (Joule's 
Equivalent) converts it to heat units to correspond with the 
other values. It should be noted that P in this expression 
stands for pressure in pounds per square foot. 

The total latent heat of vaporization is generally desig- 
nated by r or by L, and it follows from what has preceded 
that 

r = p+APu (13) 

The value of r for atmospheric pressure, that is, for a 
temperature of vaporization of 212° F., is very often used 
in engineering and should be memorized. Its value is now 
generally taken as 970.4 B.t.u., though recent work would 
seem to indicate a value of about 972 as nearer the truth. 

22. Total Heat of Dry Saturated Steam, X or H. The 
total heat required to convert a pound of water at 32° F. 
into a pound of saturated vapor at some temperature t v is 
called the totcl heat cf the steam or the heat above 32° and 
is designated by X or by H. It is obviously the sum of the 
quantities which have just been considered, so that 

\ = q+r = q+p+APu. . . . . (14) 

23. Total Heat of Wet Steam. In practical work the 
engineer seldom deals with pure saturated steam, the satu- 
rated vapor nearly always carrying in suspension more or 
less liquid water at its own temperature. To distinguish 
between saturated steam which carries liquid water and that 
which does not, the former is called wet steam or wet 
saturated steam, and the latter dry saturated steam. 

The condition of dryness or wetness is described by what 
is known as the quality of the steam. Dry saturated steam 
is said to -have a quality of 100 per cent while saturated 
steam carrying 10 per cent by weight of liquid is said to 



34 STEAM POWER 

have a quality of 90 per cent. Quality expressed as a 
decimal fraction is designated by the letter x, so that if x 
is said to be equal to 0.8 in referring to a certain sample of 
steam, it means that that steam sample consists of 80 per 
cent by weight of saturated steam and 20 per cent liquid 
at the same temperature. 

Since the water in wet steam has the same temperature as 
the steam, it contains all the heat of the liquid which it would 
contain if it had been converted into steam, but it obviously 
contains no latent heat of vaporization. It follows that the 
total heat in a pound of wet steam (one pound of a mixture of 
saturated steam and water) with quality equal to x is 

Heat per pound = q-\-xr = q-\-xp-{-xAPu . . (15) 

The letter X should never be used in designating the total 
heat per pound of wet steam, as it has been chosen as the 
symbol of the total heat per pound of dry, saturated steam. 

24. Heat of Superheat. When the temperature of 
saturated steam is raised by the addition of more heat, that 
is, when it is superheated, a very definite quantity of heat 
is required. The quantity required per pound per degreee 
would, by definition, be the specific heat of the material in 
question. 

If the specific heat of superheated steam were reasonably 
constant, the heat required to raise its temperature at 
constant pressure from saturation temperature to some 
higher value fe would be given by the expression 

Heat required per pound = C P fe — U) 

but superheated steam, as handled by the engineer, is 
generally comparatively near the saturated condition, and 
under these circumstances the values of the specific heat vary 
rapidly with changes of pressure and temperature. The 
extent of these variations is shown in Fig. 13. It will be 
observed that for low pressures the specific heat is approxi- 



STEAM 



35 



mately constant at a value below 0.5 for any given pressure, 
but that for very high pressures it varies widely over a 
comparatively small temperature range. Thus at 600 lbs. 
per square inch the specific heat changes from unity at 
about 510° F. to 0.6 at about 550° F. 

Practically, it is customary to use the type of equation 
just given and to substitute a mean specific heat over the 
required temperature range for the specific heat which can- 
not be assumed constant without too great an error. The 



1.0 



3- 9 



> 

1.7 

m 

u 



2 c 

CU .5 





















Ct/I 

'£/ 1 
w 1 










>7 \ 
c/ \ 
H' 1 \ 










§1 \ \% 


V 








<v Y5 \ 


v\ 






-^Lbs. 


per sq. ir 




__r"~~'— — «, 


. 















100 200 300 400 500 

Temperatures Deg. Fahr. 



Fig. 13. — Progressive Values of Specific Heat, C P , Water Vapor. 

equation for heat required to raise the temperature from 
U to fe is then 



Heat of superheat per pound = C pm (fe— Q. 



(16) 



in which C pm stands for the mean specific heat at constant 
pressure over the temperature range from t v to £2. 

Values of mean specific heats of superheated steam are 
given in Fig. 14, the values indicated by the curves giving 
the mean specific heat between saturation temperatures and 
various higher temperatures at different pressures. 



36 



STEAM POWER 



\ 



25. Total Heat of Superheated Steam. The total heat 
required to convert one pound of water at 32° F. into 
superheated steam at a temperature of ^° F. under constant 
pressure conditions is obviously 

Total heat per pound = g-f-r+ C pm fo — U). . (17) 



.70 



W 

o 

S-60 

o 

o 

ft 

02 



.50 - 



,.40 



v\ 


\\ 










\ 


\\ 


V 


>.<%,? 






\ \ 


\\ 


\ \ 


7^%/ 

^ ^ 








s\ 


X N& 


^^ 






^ 






^^-»^. 


C^: 




«^__ 




50 


— ■ 


i^^ 












15 


































50 



100 150 200 250 

Temperatures above Saturation*^. 



Fig. 14. — Variation of Mean Specific Heat, Water Vapor. 

and representing the degrees of superheat (fe — £») by D, as 
is customary, this becomes 



Total heat per pound = q-\-r-\-C pm D. 



(18) 



26. Specific Volume of Dry Saturated Steam, V or S. The 
volume occupied by one pound of a substance is spoken of as 
the specific volume of that material. In the case of dry 
saturated steam there are as many specific volumes as there 
are pressures under which the steam can exist. These 
values are generally tabulated in steam tables and are 
represented by the letter V or the letter S. 



STEAM 



37 



The values of the specific volumes of steam at different 
pressures are given in Fig. 15. It is important to note the 
very gradual change of specific volume at high pressures 
and the very rapid change and enormous increase at low 
pressures. These facts have considerable influence on steam 
engineering practice. 



DZV 


M 












































.480 








































































































































440 














































£400 

a 


























































































|360 

p 


























































































^320 


























































































.§280 

g 


























































































£240 






















































































2 

3 200 
to 

CO 






\ 












































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p: 160 






\ 












































\ 


v 






































120 








\ 














































\ 














































\ 


\ 


































.80 














\ 


V. 






























40 









































































































































2 4 (3 8 10 12 14 16 18 20 22 
Specific Volume of Dry Saturated Steam (Cubic Feet) 

Fig. 15. — Pressure-Volume Relations, Saturated Water Vapor. 

A curve giving properties of saturated steam is called a 
saturation curve, so that this name may be, and often is, 
applied to the curve given in Fig. 15. 

The volume occupied at any pressure by half a pound 
of dry saturated steam will obviously be half that occupied 
by one pound of such material at the same pressure, and 



38 STEAM POWER 

the same statement can be made for any other fraction of 
a pound. It follows that if the small volume occupied by 
liquid water in wet steam be neglected, the volume occupied 
by one pound of steam (mixture) of 50 per cent quality 
can be assumed equal to half that occupied by an equal 
weight of dry saturated steam at the same pressure. A 
similar statement could of course be made for any other 
quality and a corresponding fraction. 

Hence if one pound of " wet steam " at a given pressure 
is found to have such a volume that it would be indicated by 
point b in Fig. 16, the quality of this material must be given 

by the expression x = — -if the volume occupied by the 




ac 

liquid water in the mixture be 
neglected. 

27. Specific Density of Dry 

Saturated Steam, - or 8. The 

weight per cubic foot of saturated 

steam is spoken of as its specific 

density. The specific density 

is obviously the reciprocal of 

the specific volume and is there- 
Fig. 16. — Determining Quality 

from Volume. f re — 

28. Reversal of the Phenomena Just Described. If any 

process which has resulted in the absorption of a quantity 
of heat by a substance be carried through in the reverse 
direction, the same amount of heat will be liberated. It 
follows that a pound of dry saturated steam will give up 
the total latent heat of vaporization when condensed to 
liquid at the same temperature, and that the resultant pound 
of hot water will give up the total heat of the liquid if cooled 
to 32° F. 

29. Generation of Steam in Real Steam Boiler. The 
steam boiler is equivalent to a vessel partly filled with water 



STEAM 



39 



and fitted with means for supplying heat to the water and 
for carrying off the vapor formed. This is shown diagram- 
matically in Fig. 17. At first glance this would not seem 
to be at all similar to the cylinder and piston already con- 
sidered, but it really is the exact equivalent so far as the 
generation of steam is concerned. The flow of steam out 
of the steam-pipe is restricted to the extent necessary to 
maintain a high and constant pressure within the boiler, and, 
when in regular operation, steam is formed within the 




Fig. 17. — ■Foni.ation of Steam in a Steam Boiler. 



boiler under this pressure just as fast as necessary to replace 
that flowing out. 

By picturing the steam as flowing out in layers or lamina 
these lamina can be imagined as taking the place of the 
piston in the apparatus of Fig. 11, and each pound of steam 
formed will then push a piston before it exactly as was 
assumed in the previous discussion. 

30. Gauge Pressure. The steam pressure in a boiler is 
commonly determined by means of an instrument called a 
pressure gauge. These instruments are almost always con- 
structed about as shown in Fig. 18 (a) and (6) . The Bourdon 
spring is a tube of elliptical section bent approximately into 



40 



STEAM POWER 



the arc of a circle. One end of this tube is connected directly 
to the pressure connection of the gauge and the other end 
is closed and connected to a toothed sector as shown. 

When the pressure inside a tube of this character is 
increased, the tube has a tendency to unroll or straighten 
out, and in so doing it moves the toothed sector in such a 
way as to rotate the pointer or gauge hand and make its end 
move over the scale in the direction of increasing pressure. 
With diminishing pressure the tube again rolls up and 
rotates the hand in the opposite direction. 





Treasure 
Connection 



Pressure 
Connection 



(a) (6) 

Fig. 18. — Bourdon Pressure Gauge. 



Instruments of this kind are so made and adjusted that 
the hand points to zero when the gauge is left open to 
the atmosphere. Under such conditions the pressure inside 
the tube is equal to that of the atmosphere and is not zero. 
The gauge therefore only indicates pressures above atmos- 
pheric on its scale, and the total pressure inside the boiler 
is really that shown by the gauge plus that of the atmos- 
phere. 

Pressures as indicated by the gauge are called gauge 
pressures. Pressures obtained by adding the pressure of 



STEAM 41 

the atmosphere to the reading of the gauge are known as 
absolute pressures. Then 

Absolute Pressure = Gauge Pressure -{-Atmospheric Pressure 
and 

Gauge Pressure = Absolute Pressure — Atmospheric Pressure. 

In accurate work the existing atmospheric pressure 
should be determined by means of the barometer, but for 
ordinary, approximate calculations and for cases in which 
no barometric data are available, it is customary to assume 
the pressure of the atmosphere to be equal to 14.7 lbs. per 
square inch. This is very nearly true, on the average, at 
sea level, but is generally far from true at higher elevations. 

PROBLEMS 

1. Determine by means of the steam tables the temperatures, 
total heats, heats of liquid, internal and external latent heats, and 
the specific volumes of 1 lb. of dry, saturated steam under the fol- 
lowing absolute pressures (lbs. per sq. in.) : 15, 50, 95, 180 and 400. 

2. Determine the heats of the liquid, latent heats of vapor- 
ization and total heats for 2 lbs. of dry saturated steam at the 
following temperatures in °F.: 101.83, 212 and 327.8. 

3. Determine the volumes occupied by 2 lbs. of dry saturated 
steam under the conditions of problem 2. 

4. Determine the heats of the liquid, latent heats of vapor- 
ization and total heats for 1 lb. of saturated steam with a quality 
of 90% at the following absolute pressures: 25, 50, 75, 125. 

5. Determine the total heat above 32° F. in 12 lbs. of saturated 
steam with quality of 97% at a pressure of 125 lbs. per square inch 
absolute. 

6. What space will be filled by 20 lbs. of dry saturated steam 
at a pressure of 150 lbs. per square inch absolute? 

7. What space will be filled by 20 lbs. of saturated steam at a 
pressure of 150 lbs. per square inch absolute and with a quality 
of 95% if the volume occupied by the water present be neglected? 

8. How many pounds of dry saturated steam at a pressure of 
75 lbs. per square inch absolute will be required to fill a spaceof 
10 cu. ft.? 



42 STEAM POWER 

9. How many pounds of saturated steam with quality 96% 
and at a pressure of 1 10 lbs. per square inch absolute will be required 
to fill a space of 8 cu. ft.? 

10. How much external work, measured in B.t.u., is done when 
1 lb. of water at the temperature of 212° F. is converted into dry 
saturated vapor at the same temperature? 

11. How much external work, measured in foot-pounds, is 
done when 2 lbs. of water at a temperature of 212° F. are converted 
into 90% quality steam at the same temperature? 

12. How much heat is required for doing internal work during 
the vaporization of 1 lb. of water under such conditions that the 
total latent heat of vaporization is 852.7 B.t.u. and the external 
latent heat is 83.3 B.t.u.? 

13. What is the quality of steam containing 1000 B.t.u. above 
32° F. per pound when under a pressure of 150 lbs. per square 
inch absolute? 

14. Heat is added to 1 lb. of mixed steam and water while the 
pressure is maintained constant at 100 lbs. per square inch absolute. 
The percentage of steam in the mixture is increased thereby from 
50% to 95%. 

(a) How much heat was added? 

(b) How much internal latent heat was added? 

(c) How much external latent heat was added? 

15. How much heat is required to completely vaporize 1000 
lbs. of water at a temperature of 92° F. when pumped into a boiler 
in which steam is generated at a pressure of 150 lbs. per square 
inch gauge? Note that heat above 32° F. in 92° F. water is given 
as q in steam tables for a temperature of 92° F. 

16. Find the amount of heat necessary to produce in a boiler 
200 lbs. of steam having a quality of 97% at a pressure of 100 lbs. 
gauge when the feed water has a temperature of 205° F. 

17. What volume would be occupied by the material leaving 
the boiler in problem 16, neglecting volume occupied by water? 



CHAPTER IV 

THE IDEAL STEAM ENGINE 

31. The Engine. If the cylinder and piston assumed in 
the discussion of the last chapter be imagined as turned 
into a horizontal position and fitted with a frame, piston 



-Fly wheel 




Fig. 19. — Simple Steam Engine. 

rod, crosshead, connecting rod, crank shaft and flywheel 
as in Fig. 19, a device results which might be used as a 
steam engine for the production of power. By adding heat 
to, and taking heat from, the water and steam in the cylin- 
der in the proper way and at the proper time, the water 
and steam, or working substance, can be made to do work 
upon the piston. The piston can transmit this work 
through the mechanism to the rim of the flywheel, and it 
can be taken from the rim by a belt connected to a pulley 
on a machine which is to be driven, 

43 



44 



STEAM POWER 



To make the analysis easier, a simplified type of engine 
will be assumed. It is shown in Fig. 20 and consists of the 
same cylinder, piston and piston rod as just described. 
A wire is fastened to the end of the piston rod and run back 
over a pulley in such a way that a weight fastened to the 
free end of the wire will be raised if the piston moves out. 
The weight is made up of two parts, one large and one small. 
When both are on the wire the pull which they exert causes 
the piston to exert a high pressure upon whatever is con- 



Cylinder 




Volume 

Fig. 20. — Simplified Steam Engine. 



tained in the cylinder. When only the small weight hangs 
on the wire, the piston exerts a much lower pressure upon 
the material in the cylinder. 

Imagine that the piston and the walls of the cylinder 
are made of some ideal material which will not receive or 
conduct heat. Imagine also that the cylinder is fitted 
with a permanent head which is a perfect conductor of 
heat. These conditions are of course ideal but are assumed 
for the sake of simplicity. 

Assume further that, when one pound of water is con- 



THE IDEAL STEAM ENGINE 45 

tained in the cylinder and the piston is driven into the 
cylinder by the two weights until the space between the 
piston and the cylinder head is just large enough to 
contain the pound of water, the piston exerts a high 
pressure equal to Pi pounds per square foot against the 
water. The volume of this water and the pressure upon 
it can be represented by the point a of the PV diagram, 
Fig. 20. 

32. Operation of the Engine. With conditions as de- 
scribed in the preceding paragraphs, imagine a flame or 
other source of heat at high temperature to be brought into 
contact with the conducting cylinder head and to pass heat 
into the cylinder, raise the temperature of the water within 
to the temperature of vaporization and ultimately vaporize 
it. As the water vaporizes it will push the piston out of the 
cylinder just as described in the last chapter and a hori- 
zontal line such as ab in Fig. 20 will represent the increase 
of volume (vaporization) at constant pressure. The point 
b may be assumed to represent the volume of one pound 
of dry saturated vapor at a pressure P\. Obviously the 
steam, as it is formed, does work in driving out the piston 
against the resistance offered by the weights which must be 
raised. 

If a stop is provided which will prevent the movement 
of the piston beyond the position corresponding to the 
point b, it will be possible to remove the larger weight when 
that point is reached and the high pressure steam will hold 
the piston and rod hard against the stop. If now some 
cooling medium is applied, such as a large piece of ice held 
against the conducting head of the cylinder or water running 
over that head, heat will be abstracted and a partial con- 
densation of the steam within the cylinder will occur. 
As condensation progresses the pressure will drop because 
there will be less and less steam, by weight, in a given 
volume. Such a process, would be indicated by the line 
be which represents a drop of pressure, while the volume 



46 STEAM POWER 

contained within the cylinder walls between head and 
piston remains constant. 

When some point c is reached, the steam pressure will 
have been reduced to a value equal to that exerted by the 
small weight, and the piston will be driven in toward the 
cylinder head while the heat absorbing medium continues 
to remove heat from the steam and to cause further conden- 
sation. The combination of piston motion and heat ab- 
sorption will be so regulated that the pressure remains con- 
stant at P 2 during this process, because the weight will 
move the piston inward just as fast as necessary to main- 
tain a constant pressure. If sufficient heat is absorbed, 
the pound of material within the cylinder will ultimately 
all be condensed or liquefied and will just fill the volume Va. 

The heat absorbing body may now be removed and an 
infinitesimal motion of the piston toward the head would 
serve to raise the pressure on the liquid water from P 2 to 
Pi so that the volume Va may be taken equal to the volume 
V a and the line da may be assumed to be vertical. It 
would then represent an increase of pressure at constant 
volume. This might be caused by hanging a weight of the 
larger size on the wire when condition d was reached. 

Having brought the material, or working substance, 
back to the conditions originally shown at a, the high 
temperature source of heat can again be brought in contact 
with the end of the cylinder and the entire cycle carried 
through once more. There is obviously no reason why it 
could not be repeated as often as desired. 

33. Work Done by the Engine. If the device just 
described is to serve as a steam engine, it must actually 
make mechanical energy available, that is, it must convert 
into mechanical form some of the heat energy supplied it. 
It is now necessary to see whether it does so. 

Water vaporizing and increasing in volume as from 
V a to Vb was shown in the last chapter to do work upon the 
piston confining it. Work has been shown to be equal to 



THE IDEAL STEAM ENGINE 47 

(total force X total distance) and in this case if L repre- 
sents the distance in feet traveled by the piston, the work 
done by the steam upon the piston while the latter moves 
from a' to b' must be 

Work done on piston = total force X distance 

= PiXarea of piston XL . . . ft.-lbs. 

But the product of area of piston in square feet by 
distance traveled in feet is equal to the piston displace- 
ment or volume swept through by the piston, that is 
(Vb— Va) cubic feet. Therefore 

Work done on piston = Pi (V b - V a ) ft.-lbs. . (19a) 

Pl(V b -V a ) 



778 



B.t.u. . (196) 



The first form of this expression Pi(Vb—V a ) is very 
obviously represented by the area under the line ab in 
Fig. 20 and this area therefore represents the work done by 
the steam upon the piston during the change of volume at 
constant pressure represented by that line. While the 
steam is supplying this amount of energy to the piston or 
doing this amount of work upon the piston, the latter does 
an equivalent amount of work upon the weights if friction- 
less mechanism be assumed. In such a case the total 
weight hung on the wire multiplied by the distance raised 
would therefore give the same result in foot-pounds as 
that just obtained. 

It should be noted that Eq. (19) is merely an expression 
of the external work done during vaporization, that is, 
an expression of the amount of heat which is used for the 
doing of external work. It is the exact equivalent of the 
external latent heat previously discussed. In fact, the 
group of symbols APu is really a condensation of Eq. (19) 

formed by putting A for ^^ and u for (V b — V a ). 



48 STEAM POWER 

The line cd also represents a change of volume at con- 
stant pressure and the same type of formula as applied 
to ab will express the work done during this process. In 
this case, however, the piston is being pushed into the 
cylinder by the small weight against the pressure of the 
steam, and energy is being supplied to push the piston in. 
This energy is equal to the weight of the small weight 
(pounds) multiplied by the distance it falls (feet). The 
piston is therefore doing work upon the steam, and the 
amount is 

Work done on steam = P 2 (V e -V d ) ft. -lbs. . . (20) 

P2(Vc-V d ) 



778 



B.t.u. . . (21) 



The first form of expression also represents the area under 
the line cd and this area therefore represents the work 
done by the piston upon the steam mixture in the cylinder 
during the process represented by cd. 

No work can be done by steam on piston or by pistoi? 
on steam during the processes represented by be or da 
because both the weights and the piston are stationary 
during these changes and it has already been shown that 
work involves motion. 

The total work done upon the piston by the steam 
is therefore represented by the area abef and this amount 
of energy is used in raising the two weights through a 
vertical distance equal to the piston travel. Some of 
this energy, or its equivalent, will have to be returned an 
instant later, however, in order that the piston may do the 
work shown by the area cdfe upon the steam. It is returned 
by the small weight dropping through a distance equal to 
the travel of the piston. The net mechanical energy 
made available by carrying through the series of processes 
is therefore represented by the area (abef) — (cdfe) = (abed) 
or the area enclosed by the four lines representing the 



THE IDEAL STEAM ENGINE 49 

pressure and volume changes experienced by the working 
substance during one cycle of events. It is equal to the 
work done in raising the larger weight a vertical distance 
equal to the travel of the piston. 

This net energy made available is obviously 

Energy made- available = Pi ( V b — V a ) — P2 ( V c — Va) 

= (Pi-P 2 )(n-K)ft.-lbs. (22) 

(Pi-P 2 )0V-F.) 



778 



B.t.u. (23) 



Since this amount of energy is made available while one 
cycle of events is being carried out and since the cycle 
can be repeated time after time if sufficient heating and 
cooling mediums are available, any quantity of mechanical 
energy can be produced from heat energy by repeating the 
cycle a sufficient number of times. This would correspond 
to picking up a number of the larger weights which were 
slid on to the wire at the lower elevation and slid off at 
the higher. 

This repetition of cycles would correspond, in a real 
engine, to running at such a speed that the required number 
of cycles would be produced in a given time to make avail- 
able the amount of mechanical energy required. 

Or, the power made available per cycle could be increased. 
This is easily seen by an inspection of Eq. (22). Increas- 
ing the value of either of the right-hand terms will obviously 
increase the amount of energy made available. The 
value of (P1 — P2) can be increased by raising the initial 
pressure Pi or by lowering the final pressure P2. The value 
of (Vb— Va) may be increased by using more than one pound 
of material, thus increasing both the volume Vb of the satu- 
rated steam formed and increasing the volume V a of the 
liquid water, but getting a greater numerical value for 
(Vb— Va). This would correspond in a real case to using 
a larger cylinder and therefore a larger engine. 



50 STEAM POWER 

ILLUSTRATIVE PROBLEM 

An engine of the type described is to work with a maximum 
pressure of 100 lbs. per square inch absolute and a minimum 
pressure of 15 lbs. per square inch absolute. The cylinder is to 
be of such size that 1 lb. of water is used and the steam is to be 
dry and saturated at the point b of the cycle. 

Find: (a) the amount of mechanical energy made available 
per cycle; (b) the amount of energy made available per minute 
if 150 cycles are produced per minute; and (c) the horse power of 
the engine. 

It will first be necessary to find the piston displacement required 
and the space necessary between piston and cylinder head to 
accommodate the pound of water in liquid form. The steam tables 
give the volume of one pound of dry saturated steam at 100 lbs. 
per square inch as 4.429 cu.ft. and the volume of one pound of 
water may be taken as 0.017 cu.ft. The values of the various 
volumes and pressures will therefore be 

Va = V d = 0.017 cu.ft,; 

V b = Vc= 4.429 cuit.; 

p a =p b = 100X 144 = 14,400 lbs. per sq.ft.; 

p c =p d = 15X144 =2160 lbs. per sq.ft. 

(a) Using Eq. (22) the amount of mechanical energy made 
available per cycle will be 

(Pj -P 2 )(V b - V a ) = (14,400 -2160) (4.429 -0.017) 
= 12,240X4.412; 
= 54,002.88 ft.lbs. 

(6) If 150 cycles are produced per minute, the total amount 
of mechanical energy made available per minute must be 

150X54,002.88 =8,100,300 ft.-lbs. 

(c) The horse power must then be 

, 8,100,300 

h - p - = ^ooo- =245+ - 

34. Heat Quantities Involved. It is a very simple 
matter to determine the quantity of heat which must be 
supplied to produce the process ab, and the quantities of 



THE IDEAL STEAM ENGINE 51 

heat which must be removed to produce the processes 
be and cd, This can be done by making use of the known 
properties of water and steam as given in the steam 
tables. 

The water at d must be at the temperature of vaporiza- 
tion corresponding to pressure P2 since it has just been 
formed by condensation from steam under that pressure. 
It therefore contains the heat of the liquid corresponding 
to that pressure. If it is to be vaporized at pressure Pi, it 
must first be raised to the higher temperature corresponding 
to that pressure. The amount of heat required to do this 
will obviously be the difference between the heat of the 
liquid at the temperature corresponding to Pi and the heat 
of the liquid at the temperature corresponding to P2, 
These can be found in the steam tables. 

The latent heat of vaporization at Pi must then be added 
to cause the increase of volume shown by ab. This can 
also be found in the steam tables for any given case. 

The quantity of heat which must be removed to produce 
the processes represented by be and cd can be found sim- 
ilarly from steam table values, although the exact method 
of procedure is not quite as obvious as in the preceding 
cases. 

Assuming that it is possible to find the heat supplied, 
Qi, and the heat removed, Q2, it is obvious that the energy 
made available in mechanical form, per cycle, must be 
equal to (Q1 — Q2) B.t.u., since this is the amount of heat 
energy which has disappeared and since it cannot have 
been destroyed. This may be put in the form of an equa- 
tion, thus 

Energy made available = Qi — Q2. . . (24) 

If the proper substitutions are made in this formula and it 
is then simplified, it becomes 

Energy made available = (APu) Pl — x c (APu) P2 B.t.u., (25) 



52 STEAM POWER 

in which 

(APa) Pl = the external latent heat at pressure Pi; 
(APu)p 2 = the external latent heat at pressure P2, and 
x c = quality at point c, which can be found from 
the ratio of dc to dc'\ 

Numerical substitution in this equation for any given 
case will show that it gives exactly the same values as would 
be obtained by the use of Eq. (23). 

It is to be noted particularly that the energy made 
available is actually less than the external latent heat at 
the higher pressure, while the heat supplied must be equal 
to the total latent heat plus some of the heat of the liquid. 
An inspection of the steam tables will show that the exter- 
nal latent heat for ordinary steam pressures forms a very 
small fraction of even the total latent heat, and therefore 
the mechanical energy made available for a given expendi- 
ture of heat energy is very small in the case under dis- 
cussion. 

35. Efficiency. The term efficiency is used in engineer- 
ing as a measure of the return obtained for a given expendi- 
ture. It may be defined in any one of the following ways: 

^^ . Useful result 

Efficiency = 



Expenditure made to obtain that result 
Result 



Effort 

Output 
Input 



(26) 



In the case of a heat engine, the useful result is the 
mechanical energy obtained by the operation of the engine, 
while the expenditure made is the heat which is supplied. 
For this case efficiency may therefore be denned by the 
expression 



THE IDEAL STEAM ENGINE 53 

_ . ^ . Mechanical energy obtained per cycle 

Engine efficiency = ^— ^—z f 

Heat supplied per cycle 

=£ ™ 

=^ • • • • <*> 

in which 

E stands for mechanical energy obtained, 
Qi stands for heat supplied, and 
Q2 stands for heat rejected. 

In the case of the type of steam engine just considered, 
this efficiency would have a value between 6 and 8 per 
cent for ordinary pressures. That is, the engine would 
produce in mechanical form only 6 to 8 per cent of the 
energy supplied it in the form of high temperature heat. 
Moreover, these figures would hold only for a theoretically 
perfect engine; a real engine built to operate upon this 
cycle would probably give efficiencies of the order of 2 to 
3 per cent. The reasons for this great discrepancy will be 
discussed in a later chapter. 

36. Effect of Wet Steam. In what has preceded, it 
was assumed that the pound of steam was completely vapor- 
ized along the line ab so that dry, saturated steam existed 
in the cylinder at b. It might, however, be assumed that 
vaporization was incomplete at the upper right-hand 
corner of the cycle, so that this corner occurred at a point 
to the left of b in Fig. 20 and with a quality x less than 
unity. 

Under such conditions, the cylinder would not have to 
be so big, since the maximum volume attained by the steam 
would be smaller than in the preceding case. The work 
done per cycle would obviously be smaller in quantity, 
because the area enclosed within the lines of the cycle would 
be smaller. It can also be shown that the efficiency would be 
lowered by incomplete vaporization, 



54 STEAM POWER 

37. Application to a Real Engine, The engine which 
has been described in the preceding paragraphs could easily 
be converted into the counterpart of a real engine by sub- 
stituting connecting rod, crank shaft and flywheel for wire, 
pulley and weights as described in the first paragraph of 
this chapter. It could then be made to do work in just 
the same way as has been described; some of the energy 
made available during the outstroke would be used for 
overcoming resistance at the shaft, that is, doing useful work, 
and some of it would be stored in the flywheel which would 
speed up slightly. The energy which must be expended on 
the steam during the return stroke would be obtained by 
allowing the flywheel to slow down and thus deliver suf- 
ficient kinetic energy to drive the piston back against the 
low-pressure steam. The cycle and the efficiency would 
thus, theoretically, be exactly the same as those just in- 
vestigated. 

Great difficulty would, however, be met in a real engine 
if the steam had to be formed and condensed within the 
cylinder, and another method which gives the same results 
is therefore used. Steam is generated in a boiler and 
allowed to flow into the cylinder and push out the piston 
just as though it were actually being formed in the cylin- 
der as previously described. When the piston reaches the 
end of its outstroke the inlet valve is closed and the exhaust 
valve is opened, allowing some of the steam to blow out 
into a space in which a lower pressure exists. As the 
piston stands still at the end of its stroke while the pres- 
sure drops, the line be is produced as in the previous descrip- 
tion, but by a different method. The piston then returns 
and drives the remaining steam out of the cylinder at a 
constant pressure theoretically equal to that of the space 
into which the steam is being forced or exhausted. The 
line cd is thus produced and the closure of the exhaust 
valve and opening of the admission valve when d is reached 
will start the cycle over again. 



THE IDEAL STEAM ENGINE 55 

In order to get more work out of a given size of cylinder 
and to obviate the necessity of giving back energy which 
has already been given out, engines are generally made to 
take steam on both sides of the piston. They are then 
known as double acting engines. In this case the steam 
admitted on one side of the piston would supply the energy 
necessary both for overcoming the resistance due to the 
load and for driving out the low-pressure steam on the 
other side of the piston. On the return stroke conditions 
would be just reversed. 

38. Desirability of Other Cycles. The cycle of opera- 
tions described in preceding paragraphs is the most inef- 
ficient of all those actually used, that is, it gives the small- 
est return for a given amount of heat supplied. This 
is because only the external latent heat supplied is con- 
verted into mechanical energy and part of that energy 
must be returned to complete the cycle. All of the heat of 
the liquid as well as all the internal latent heat supplied along 
ab passes through the engine without conversion and is ex- 
hausted. 

Therefore, cycles which differ from that described 
in such a way as to make it possible to convert into mechani- 
cal energy some of the internal latent heat and possibly 
some of the heat of the liquid should be highly desirable 
as they ought to yield a larger return of mechanical energy 
for the same total amount of heat supplied. Two such 
cycles are commonly used; they may be described as the 
Complete-expansion cycle and the Incomplete-expansion cycle. 
The former is used in steam turbines, the latter in most 
reciprocating steam engines. The rectangular cycle which 
has just been described is used in duplex pumps and similar 
apparatus. 

39. The Complete-expansion Cycle. This cycle, which 
is also known as the Clausius and as the Rankine cycle, 
starts just the same as that already described. This is 
shown in Fig. 21. The pressure on, say, a pound of water 



56 



STEAM POWER 



is raised from P2 to Pi and its temperature is raised from 
that of vaporization at P2 to that of vaporization at Pi. 
After this it is vaporized, giving the increase of volume 



Coiled Spring 


b 


\c[ 




t h K'h A 




\ A *_ A 

\ /' \ / \ / \ }, 1 


2 

1 


}&i j 




J Volume 

p' | 6 ' ic 




£.- ■ 






s 

I 

p 2 


« 


1 
1 
6 

/V/sS Saturation Curve 

/v/y/AS. \ Work done by Steam 
///w/W)v S /. ( luriag expansion 6-G 


d 


WMMiMMk 



Volume 
Fig. 21. — Complete Expansion Cycle or Clausius Cycle. 

shown by ab. The supply of heat is then stopped. The 
cylinder of the engine is made larger than in the preceding 
type so that when the point b' is reached the piston can 
travel still further, and it is allowed to do so, that is, the 



THE IDEAL STEAM ENGINE 57 

high-pressure steam is allowed to push it further out. This 
can be pictured by imagining the steam to act like the 
compressed spring shown in the figure and to push the 
piston in much the same way as does the spring. The line 
bici shows the decreasing pressure exerted on the piston by 
the spring as the latter expands so as to get longer and 
longer. Because of the properties of a spring this is a 
straight line. The line be shows the decreasing pressure 
exerted on the piston by the steam as the latter expands 
so as to occupy greater and greater volumes. Because of 
the properties of steam this line is curved instead of straight. 

Work will be done on the piston by the expanding 
steam during the process be and the amount of this work 
will be indicated by the area under the line be as shown 
in the figure. This work must have been done by the 
expenditure of energy on the part of the steam and since 
no energy was added after the point b was reached the work 
must have been done at the expense of heat energy contained 
in the steam at b. It has already been shown that the 
heat above 32° in the steam at b is equal to the sum of the 
heat of the liquid and the internal latent heat, and some 
of this heat must obviously be used for the doing of work 
along be instead of being entirely rejected to the cooling 
medium as in the preceding cycle without " expansion." 

The expansion of the steam continues until the " back 
pressure " P2 is reached. The cooling medium may then 
be imagined to be brought into use and to abstract such 
heat of vaporization as may remain in the steam besides 
absorbing the equivalent of the work done on the steam 
by the returning piston, thus giving the process shown by 
the line cd. 

If the expansion line be of the cycle just described 
could be carried out within walls constructed of such mate- 
rial that it would not give heat to nor take heat from 
the steam, it is obvious that any heat energy lost by the 
steam during the expansion could be lost only by conver- 



58 STEAM POWER 

sion into mechanical energy. An expansion of this kind 
is called an adiabatic expansion. 

In the figure, the curve of adiabatic expansion is shown 
in its correct position with respect to the saturation curve 
and it is obvious that for an adiabatic expansion, starting 
with dry, saturated steam, the quality decreases as the expan- 
sion progresses. 

Comparison with- Cycle without Expansion. The heat 
supplied is the same in both of the cycles just considered 
when they operate between the same two pressures, but 
the mechanical energy obtained in the case of the complete 
expansion cycle is much greater. In Fig. 21, for instance, 
the mechanical energy obtainable with the cycle first 
described is represented by the area abd'd while that obtain- 
able with the complete expansion cycle with the same 
heat supply Qi is represented by the same area abd'd plus 
the additional area bed' . The efficiency of the complete 
expansion cycle is therefore very much higher than that 
of the cycle without expansion. 

For conditions similar to those giving a theoretical 
efficiency of about 6 per cent without expansion, the com- 
plete expansion cycle will give a theoretical efficiency of 
about 12 per cent and this figure can be doubled by 
expedients which will be considered later. 

The cylinder required for the production of the com- 
plete expansion cycle would be much larger than that re- 
quired for the other cycle if both used the same weight of 
steam per cycle. The proportion would be in the ratio 
of the volume shown at c in Fig. 21 to the volume shown 
at b. But the complete expansion cycle would make avail- 
able much more energy per pound of steam than would 
the other, so that the difference in the size of cylinders 
would not be so great if both were required to make avail- 
able the same amount of mechanical energy per cycle. 

40. The Incomplete-expansion Cycle. The shape of 
this cycle is shown in Fig. 22. It is just like the complete 



THE IDEAL STEAM ENGINE 



59 




Fig. 22. — Incomplete Expansion Cycle. 



expansion cycle down to the point c. The cylinder in which 

it is produced has a smaller volume than that used for the 

complete expansion cycle so that the piston arrives at the 

end of its stroke before it 

has opened up volume 

enough to enable the 

steam to expand all the 

way down to the lowest 

pressure (terminal or back 

pressure) . When the point 

c is reached in the real 

engine, the exhaust valve 

is opened and enough 

steam then blows out to 

reduce the pressure to the back pressure P a . The piston 

then returns and drives out the remainder of the steam as 

shown by the line de. 

In the ideal method assumed in the preceding treat- 
ment, the heat absorbing medium would be brought into 
use at c, absorbing sufficient heat to reduce the pressure 
from P c to Pa while the piston remained stationary at the 
end of its stroke. The latent heat of vaporization remain- 
ing in the steam at d would then be absorbed as the piston 
was driven back from d to e. 

Comparison with Other Cycles. The incomplete expan- 
sion cycle is intermediate between the two previously dis- 
cussed. This can be appreciated readily by an inspection 
of Fig. 22. In this figure the area abd'e represents the 
mechanical energy obtainable with the cycle without 
expansion; the area abc'e represents the energy obtainable 
from the same quantity of steam with complete expansion; 
and the area abcde represents the energy obtainable from 
ihe same amount of steam with incomplete expansion. 

The later the point at which the exhaust valve is opened, 
point c, the more nearly do efficiency and energy obtain- 
able approach the values for the complete expansion cycle. 



60 STEAM '.POWER 

The earlier the point at which the exhaust valve is opened, 
the more nearly do efficiency and energy obtainable approach 
the values for no expansion. 

Despite the lower efficiency of the incomplete expan- 
sion cycle as brought out in connection with Fig. 22 it is 
universally used on all reciprocating engines excepting 
those which make do pretense to economy and use no 
expansion. The less efficient cycle is used for the simple 
reason that complete expansion in a reciprocating engine 
does not pay commercially. For complete expansion the 
cylinder must be larger in the ratio of V c to V c > as shown 
in Fig. 22 and the work obtained by completing the expan- 
sion is a very small part of the total. In most cases it 
would not be great enough to overcome the friction of the 
engine, not to mention paying interest on the necessarily 
higher cost of the larger cylinder and accompanying parts. 

It will be shown in a later chapter that the steam tur- 
bine can economically expand the steam completely and 
the complete expansion cycle is therefore used with such 
prime movers. 



CHAPTER V 
ENTROPY DIAGRAM 

41. Definitions. In Chapter III temperature, pressure 
and volume were discussed as criteria determining the con- 
dition of water and steam. Other things may be used in 
determining the condition of such materials. One which is 
particularly useful from an engineering standpoint is known 
as entropy and is designated by the Greek letter cj>. 

For every condition of water and steam, there is a char- 
acteristic value of entropy just as there is a characteristic 
value of temperature, pressure, volume, heat above 32° F., 
etc. These values of entropy are given in the steam tables 
in just the same way as the value of temperature, pressure, 
volume, heat above 32° F., and such, are given. 

The entropy of the liquid given for any particular pres- 
sure is the change of entropy experienced by one pound 
of the liquid when its temperature is raised from 32° F. 
to the temperature of vaporization corresponding to that 
particular pressure. It might be spoken of as the entropy 
of the liquid above 32° F., just as q is spoken of as the heat 
of the liquid above 32° F. It is represented by <f>i. 

The entropy of vaporization given for any particular 
pressure is the change of entropy experienced by one pound 
of the material while changing from water at the tempera- 
ture of vaporization to dry saturated steam at constant 
pressure. It corresponds to the latent heat of vaporiza- 
tion and is designated by fa. 

The entropy of dry saturated steam at any pressure is 
the sum of fa and fa and therefore is the total change of 
entropy experienced by a pound of material in changing 

61 



62 



STEAM POWER 



from water at 32° F. to dry saturated steam at the particu- 
lar pressure in question. 

The entropy of superheat at any pressure and tempera- 
ture is the change of entropy experienced by a pound of 
dry, saturated steam at that pressure when superheated 
to that particular temperature. It is designated by <j> s . 

The entropy of superheated steam at any pressure and 
temperature is the total change of entropy experienced by 
one pound of material when changed from water at 32° F. 




•So/ I 

*-} i,/ i Wet Steam Region i o/ 

2W- 



•£/ JRegion of Incomplete! Vy^ 
' | Evaporation 



Entropy Scale 




(«) 



Fig. 23.— Temperature-Entropy Diagrams. 



to superheated steam at the pressure and temperature 
in question. It is equal to <fr+# +<^. 

42. Temperature-Entropy Chart for Steam. Entropy 
is particularly useful to the engineer because it enables him 
to draw charts which lend themselves readily to an easy, 
graphical solution of certain problems which would other- 
wise involve complex calculations. One of these charts 
is known as the Temperature-Entropy Chart. 

In making this chart, absolute temperature is generally 
plotted on the vertical and entropy above some datum tem- 
perature on the horizontal, as shown in Fig. 23 (a) and (6), 
which represents the construction of a temperature entropy 
diagram for water and steam. The entropy values on 



ENTROPY DIAGRAM 63 

this chart are plotted above 32° F. as datum tempera- 
ture. 

The water line or water curve is obtained by picking 
out of the steam tables the values of <fr, entropy of the liquid, 
for different pressures and plotting them against the abso- 
lute temperatures corresponding to those pressures. Ob- 
viously, zero of entropy will occur at the absolute tempera- 
ture corresponding to 32° F., i.e., about 492° F. abs. 

The saturation curve or dry steam curve is obtained by 
picking out of the steam tables the values of <&+<£» for 
different pressures and plotting against corresponding 
absolute temperatures. 

The entropy of vaporization is obviously shown for each 
different temperature (or pressure) by the distance between 
the water curve and the saturation curve, since the former 
is distant from the vertical axis by an amount equal to fa, 
while the latter is distant an amount equal to fa-\-fa. 

Superheating lines are drawn by picking from the steam 
tables the values of entropy above 32° F. for steam super- 
heated to different temperatures at one particular pressure 
and plotting against the proper temperatures. There will 
be as many superheating lines on the diagram as one chooses 
pressures for which to plot them. Only one is shown 
in the figure. 

One very useful property of this diagram follows from the 
fact that points on its surface indicate the condition of the 
material. For instance, if the temperature-entropy, or 
T— fa values of the material at a given condition should 
plot to the left of the liquid line, the material must be in the 
liquid condition; if they plot between the liquid line and the 
saturation curve, the material must be a mixture of liquid 
and saturated vapor; if they plot on the saturation curve, 
the material must be dry, saturated steam; and if they 
plot to the right of the saturation curve, the material must 
be superheated steam. This all follows directly from the 
definition of entropy above 32° F., as plotted in these dia- 



64 STEAM POWER 

grams. The various regions, or fields, into which the dia- 
gram divides in this way are shown in Fig. 23 (a). 

Another very useful property of this diagram follows 
from the fact that area represents heat just as area on a 
pressure-volume diagram was found to represent work. 
Thus the area under the line ab, for instance, represents 
the heat required to raise the temperature of one pound 
of water from 32° F. to the temperature at b. Similarly 
the area under the line be represents the heat required to 
change a pound of water at the temperature at b to a pound 
of dry, saturated steam at the same temperature. The 
heat required to superheat this pound of saturated steam 
at constant pressure up to the temperature shown at d 
is similarly represented by the area under the line cd. 

In this connection, it should be noted that this diagram 
is plotted above absolute zero of temperature just as the 
pressure-volume diagram is plotted above absolute zero of 
pressure. The areas in question therefore extend down 
to the absolute zero of temperature. In order to indicate 
this in Fig. 23 (6), a large part of the chart is supposed to 
have been broken out, so that the lower end of the diagram 
could be moved up into view. In Fig. 23 (a), the bottom 
of the diagram is drawn a few degrees below 32° F. and 
this is indicated by putting T>0 opposite the horizontal 
axis. 

The various areas hatched in Fig. 23 (b) indicate the 
various quantities of heat previously discussed. It should 
be understood that the areas represent the heat quantities 
only for the particular pressure which corresponds to the 
temperature indicated by T v . For a higher pressure, the 
line be would be higher and the areas proportionately 
larger; for a lower pressure the line be would be lower and 
the areas smaller. 



ENTROPY DIAGRAM 



65 



ILLUSTRATIVE PROBLEM 

Starting with liquid at a temperature Tx corresponding to the 
temperature of vaporization at a pressure of 50 lbs. per square 
inch absolute, assume the liquid raised to the temperature of 
vaporization at a pressure of 100 lbs. per square inch absolute 
and then completely vaporized. Determine the various changes 
of entropy and indicate them on a T^-chart. 

The steam tables give entropy of the liquid, 4>i, as equal to 
0.4113 for water about to vaporize under 50 lbs. per sq. in. 
absolute, and 0.4743 for water about to vaporize under a pres- 
sure of 100 lbs. per sq. in. absolute. The difference, that is, 
0.4743-0.4113=0.0630, must be 
the entropy change experienced 
by the liquid when its tempera- 
ture is raised from the lower to 
the higher value. These values 
hie shown in Fig. 24. 

The steam tables give entropy 
of vaporization, <b v , at 100 lbs. per 
square inch absolute as 1.1277. 
Adding this to the entropy above 
32° F. of the liquid at vaporiza- 
tion temperature under 100 lbs. 
pressure gives 0.4743+1.1277 = 
1.602 as the entropy above 32° 
of dry, saturated steam at 100 

lbs. per square inch absolute. These values are all indicated in 
their proper position in Fig. 24. 

The total change of entropy experienced by the material in 
changing from water at the temperature of vaporization under 
50 lbs. pressure to dry, saturated steam at 100 lbs. pressure is 
obviously equal to 0.0630+1.1277 = 1.1907. 

43. Quality from T^-chart. The entropy change ex- 
perienced by steam in the process of vaporization is directly 
proportional to the addition of heat. Thus, when half 
the latent heat has been added to one pound of material, 
the entropy change is i<j> v . In general, if a fraction x of 
the latent heat has been added, the entropy change has 
been x<f>„ during the process. Therefore, if the temperature 
entropy condition of a pound of material should plot at a 




Fig. 24. 



66 



STEAM POWER 



point such as c in Fig. 25, it follows that the material is 

a mixture of water and steam and that a fraction of the 

be 
pound equal to — is steam, the rest being water. But, 

be 
by definition, the fraction — is x, the quality of the 

material. 

The temperature-entropy chart is very useful when used 
in connection with this property of showing quality. Thus, 
in Fig. 25, the area under be, down to absolute zero tem- 
perature, represents the fraction of the latent heat of 




Entropy 

Fig. 25. — Quality from Temperature- Fig. 26. — Constant Quality 
Entropy Chart. Curves. 



vaporization per pound which must be added to give a 
pound the quality x. 

For convenience in use, constant quality lines are 
generally drawn on temperature-entropy charts. Such 
lines are shown in Fig. 26. Each line is obtained by plot- 
ting the temperature entropy conditions for a given quality 
at different pressures. For this purpose, </>„ and <j>i are 
taken from the steam tables for a given pressure. The 
numerical value of <}>„ is then multiplied by the fraction re- 
presenting the chosen quality, say 0.9, and the product 
is added to 4>i, giving the total entropy above 32° F. for 
quality 0.9 at the particular pressure chosen. The same 



ENTROPY DIAGRAM 



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68 STEAM POWER 

process is repeated with the same value of the quality, 
but with different pressures, until enough points have been 
secured to make it possible to draw a smooth line through 
them. 

44. Volume from T^-chart. Since quality changes at 
any given temperature, or pressure, are accompanied by 
volume changes, it is possible to find a series of values for 
the quality of a pound of wet steam which will make that 
pound occupy the same volume at different temperatures. 
Having found the quality which will be necessary at a num- 
ber of different temperatures, the total entropy above 32° 
F. can be found for each case and these values can then be 
plotted on the TV-chart. Connecting the points so obtained 
would give what is known as a Constant Volume Line. 

Several of these constant volume lines are shown in their 
correct positions in Fig. 27. It will be observed that, 
for each volume, the quality must increase as temperature 
(and pressure) increases in order to maintain a constant 
value for the volume occupied by one pound of mixture. 

45. Heat from T^-chart. Equations for obtaining the 
total heat above 32° F. for wet and for superheated steam 
were given in an earlier chapter. By means of these 
equations, it is possible to find a succession of values for 
quality and superheat which will give a pound of material 
any chosen heat content at different pressures. If the 
corresponding values of temperature and entropy are found 
and plotted, what is known as a Constant Heat Line results. 
Several of these lines are shown in Fig. 27. 

46. The Complete T^-chart for Steam. A very com- 
plete, graphical representation of the properties of water 
and steam can be procured by combining in one diagram 
all of the lines discussed in preceding paragraphs. Such 
a diagram is generally spoken of as the T (^-diagram or the 
T4>-chart for steam. An example of such a diagram is 
given in Fig. 28. 

This chart is very useful, as it enables one to solve by 



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TEMPERATURE-ENTROPY DIAGRAM 

TO ACCOMPANY 

STEAM POWER 

C.F.HlRSHFELD AND T.C.ULBRICHT. 

(Published by John Wiley & Sons) 
Redrawn (with permission) from larger diagram in Peabody's Steam and Entropy Tables ( Wiley & Sons) 




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70 STEAM POWER 

inspection many of the most difficult problems which arise 
in the theory and practice of using steam. As an example, 
assume that it is desirable to know what will happen if 
water at the temperature of vaporization corresponding 
to about 24 lbs. per square inch absolute has its volume 
increased indefinitely at constant temperature. The initial 
condition of the water would be shown on the water curve 
of Fig. 28 at the point at which the 700° absolute temperature 
line crosses it. Increase of volume at constant temperature 
would be indicated by a horizontal line running to the right 
from this point. Obviously, vaporization will occur at 
constant pressure (because the temperature is constant) 
and the quality will change from zero to unity at which the 
saturation curve will have been reached. Further increase 
of volume can result only in the production of superheated 
steam, since the line representing the process will rui\ 
out into the superheated steam field. It is also interesting 
to note that the pressure on the material will have to be 
decreased as the volume increases in the superheated steam 
region, as is evidenced by the fact that the horizontal line 
representing the assumed process cuts lower and lower 
pressure lines as it is extended to the right in the super- 
heated field. 

Note also that the intersections of this horizontal line 
with constant volume and constant heat lines afford the 
means of determining volume and heat above 32° F. at 
different stages of the assumed process. 

PROBLEMS 

1. Determine from the steam tables the change of entropy 
experienced by one pound of water when its temperature is raised 
from 32° F. to the temperature of vaporization under a pressure 
of 100 lbs. per square inch absolute. 

2. Determine from the steam tables the entropy change experi- 
enced by one pound of water when its temperature is raised from 
32° F. to the temperature of vaporization under a pressure of 
150 lbs. per square inch absolute. 



ENTROPY DIAGRAM 71 

3. Determine the entropy change experienced by one pound of 
water when its temperature is raised from the temperature of 
vaporization corresponding to 100 lbs. per square inch to that 
corresponding to 150 lbs. per square inch by subtracting the value 
found in Prob. 1 from that found in Prob. 2. 

4. Determine the change of entropy experienced by one pound 
of material completely vaporizing at a temperature of 327.8° F. 

5. Plot a TV-chart for one pound of water. Start by plotting 
entropy of the liquid for various temperatures; then plot entropy 
of saturated steam (above 32° F.); finally draw water line, satura- 
tion line, and several lines showing change of entropy during vapor- 
ization. 

6. Determine from a T^-chart the quality which would be 
attained by one pound of steam if it experienced a change which 
carried it from the condition of dry saturated steam at 150 lbs. 
per square inch absolute to a pressure of 25 lbs. per square inch 
absolute by a process which would plot as a vertical line on the 

7>-chart. 

7. Assume a pound of mixed water and steam to have a qual- 
ity of 80% at a pressure of 200 lbs. per square inch absolute. 
Determine from the TV-chart the heat above 32° per pound of 
mixture and the volume occupied by the mixture. Determine 
also the quality attained if the pressure of the material drops to 
20 lbs. per square inch absolute at constant entropy. How does 
the heat above 32° F. change during such a process? 

8. Assume a pound of mixture as in Prob. 7, but with a 
quality of 30% at a pressure of 200 lbs. Find all quantities called 
for in that problem. 

9. Assume a pound of material as in Probs. 7 and 8 above, 
but superheated 200° at a pressure of 200 lbs. per square inch 
absolute. Determine all quantities called for in Prob. 7. 

10. Choose a point on the T^-chart at which a constant volume 
line intersects the saturation curve. Determine the change of 
quality, entropy and heat above 32° F., if the material drops 
to half pressure at constant volume. 



CHAPTER VI 

TEMPERATURE ENTROPY DIAGRAMS OF STEAM 
CYCLES 




47. Complete Expansion Cycle. This cycle was con- 
sidered in Chapter IV and the PF-diagram was given there 
as Fig. 21. The diagram of this cycle drawn to ^-co- 
ordinates is shown in Fig. 29. The same letters are used 

to represent corresponding points 
in the two diagrams. 

The entropy change during 
the heating of the liquid is 
shown by the part of the liquid 
line between d and a, and the 
heat supplied during that process 
is represented by the area below 
the line da, measuring clear 
down to the absolute zero of 
temperature. 

The entropy change during 
vaporization is represented by the line ab and the heat 
supplied during the process is shown by the total area 
under that line. 

The adiabatic expansion of the steam is represented by 
the line be, such an adiabatic change fortunately being a con- 
stant entropy process and therefore easily drawn in this 
diagram. Obviously no heat is received or removed dur- 
ing this process, as there is no area under the line be. 

The entropy change during condensation is represented 
by the line cd and the heat rejected by the working sub- 
stance during this process is represented by the area under 

that line. 

72 



Fig. 29.— 7V-diagram, Com- 
plete Expansion Cycle. 



TEMPERATURE ENTROPY DIAGRAMS 73 

48. Area of Cycle Representative of Work. It will 
be remembered that area under a line in the PF-diagram 
represents work in foot-pounds. That diagram, however, 
gives no indication of heat received or rejected and it is 
not possible to obtain any direct idea of efficiency from it. 
In this respect, the T^-diagram is much better. Area under 
the lines da and ab in Fig. 29 represents heat supplied 
the working substance. Area under the line cd represents 
heat rejected by the working substance. The difference 
between these two, or the area enclosed within the lines 
of the cycle, must therefore represent the heat converted 
into mechanical energy per cycle. 

This diagram therefore shows directly by areas the 
heat supplied, the heat rejected, and the heat converted 
into mechanical energy. Further, the ratio of the area 
representing heat converted into work, and the area repre- 
senting heat supplied must be the efficiency of the cycle. 

Remembering also that if the lines of the cycle are drawn 
upon a T0-chart such as that given in Fig. 28, all volume 
changes, heat contents and qualities at different points 
are shown without further work, it becomes evident that 
this form of representation is decidedly convenient and far 
superior to the pressure volume method. 

49. Modifications for Wet and Superheated Steam. 
The complete expansion cycle is supposed to represent an 
idealization of what happens in a real prime mover. In real 
cases, however, the steam may arrive at the prime mover wet 
or superheated and it is desirable to investigate the method 
of representing such conditions as well as their effects. 

Wet steam corresponds to incomplete vaporization, 
i.e., a quality less than unity at the upper right-hand corner 
of the cycle. This might be shown for a given case by the 
location of the point b' in Fig. 29. The cycle would then 
be ab'c'd and a smaller amount of work would be obtained 
per pound of working substance as evidenced by the smaller 
area enclosed within the lines of the cycle, 



74 



STEAM POWER 



In the case of superheated steam, superheating occurs 
at constant pressure after vaporization is complete. This 
would be shown by the location of the upper right-hand 
corner of the cycle at some point b" on the constant pressure 
line which extends out from b. The cycle is now represented 
by abb"c"d and evidently has a different shape than it 
had in the preceding cases. Obviously the area enclosed 
within the lines of the cycle is greater than it was before and 
therefore more mechanical energy is obtained per pound 
of steam. 

50. Incomplete Expansion Cycle. The only difference 
between the incomplete and complete expansion cycles is 




Fig. 30. — T<£-diagram, Incom- Fig. 31.— 7 T </>-diagram. Cycle 
plete Expansion Cycle. Without Adiabatic Expansion. 



the termination of the expansion in the former by means of 
a constant volume line. This is shown to T^-coordinates 
in Fig. 30 in which the incomplete expansion cycle is drawn 
in heavy lines over the one in which expansion continues 
to the back pressure. 

The constant volume line is seen to cut off a corner, 
thus reducing the area representing heat converted into 
work. The heat supplied in each case is measured by the 
area under the lines ea and ab. The efficiency of the cycle 
with incomplete expansion can therefore be seen to be less 
than that of the other cycle by simple inspection of the 
diagram. 

If the adiabatic expansion is terminated at a higher 



TEMPERATURE ENTROPY DIAGRAMS 75 

pressure, as by the constant volume line c"d" in Fig. 30 r 
still more of the work area is lost, but the same quantity 
of heat is supplied, and therefore the efficiency is still lower 
than when the expansion terminated at c. Obviously 
as the point at which the adiabatic expansion is terminated 
moves nearer and nearer to b as shown in Fig. 31, the cycle 
becomes less and less efficient. If the constant volume 
line starts at b, there is no adiabatic expansion and the 
cycle becomes that previously considered as having a rec- 
tangular shape in the PF-diagram. This cycle has the shape 
indicated by abed in the T^-diagram of Fig. 31. Obviously 
it is least efficient of all as was previously shown by other 
means. 

51. Effect of Temperature Range on Efficiency. It has 
already been stated (see p. 26) that heat engines receive 
heat at a high temperature, convert some of it into me- 
chanical form and discharge the remainder at a lower tem- 
perature. Inspection of the T^-diagram shows this very 
clearly, and, remembering that the area of the cycle measures 
the heat converted, these diagrams also show how raising 
the upper temperature (or pressure) or lowering the lower 
temperature (or pressure) will increase the efficiency. It 
can be seen readily that lowering the lower temperature 
will, however, be more effective in increasing the efficiency 
than raising the upper temperature. 

PROBLEMS 

1. Draw a complete expansion cycle to T ^-coordinates for the 
following conditions (using T^-diagram for steam to get values); 
weight of working substance, 1 lb.; initial pressure, 125 lbs. 
absolute; quality at beginning of adiabatic expansion, 100% 
back pressure, 10 lbs. absolute. 

2. Determine the following values for cycle drawn in Prob. 1 : 

(a) Entropy of liquid at beginning of vaporization ; 

(b) Entropy at beginning of adiabatic expansion; 

(c) Quality at end of adiabatic expansion; 

(d) Volume at end of adiabatic expansion ■ 

(e) Entropy at end of condensation. 



76 STEAM POWEP 

3. Show by measuring the area on TV-diagrams, the increase of 
efficiency resulting from the use of an initial pressure of 175 lbs. 
absolute and from the use of a terminal pressure of 2 lbs. absolute 
in place of the values given in Prob. 1. 

4. Compare the efficiency of a complete expansion cycle with 
conditions as in Prob. 1 with a complete expansion cycle with 
same pressures but with a temperature of 500° F. at the beginning 
of the adiabatic expansion. 

5. Draw an incomplete expansion cycle to TV-coordinates 
for the same pressures as in Prob. 1, but with adiabatic expan- 
sion ending at a pressure of 15 lbs. absolute. 

6. Compare work and efficiency of the two cycles of Probs. 
1 and 5 above. 

7. Draw a cycle without expansion for the conditions of 
Prob. 1 to TV-coordinates and compare the work area with that 
obtained in Probs. 1 and 5. 



CHAPTER VII 
THE REAL STEAM ENGINE 

52. Operation of Real Engine. In previous chapters 
the ideal steam engine was considered and several cycles 
upon which it might be operated were discussed. Real 
engines are built to operate on the same cycles, but because 
of certain practical considerations they only imperfectly 
approximate the ideal performance. 

Real engines must be built of iron and steel for practical 
reasons and these metals absorb, conduct and radiate heat 
so that certain heat interchanges between the working 
substance and engine and certain heat losses occur in 
practical operation. These were eliminated in the ideal 
case by simply assuming ideal materials not possessed 
of the characteristics of real metals. 

It is also practically impossible to generate steam 
in the cylinder of a real engine as was assumed to be done 
in the ideal case. Heat is practically obtained by the com- 
bustion of fuels, and the higher the temperature attained 
the better can the liberated heat be utilized in the genera- 
tion of steam. To subject the cylinder to such high tem- 
peratures and to control the heating and cooling as neces- 
sary to produce a number of cycles in rapid succession would 
lead to rapid wear and great practical difficulties. It has 
been found best to generate the steam in a boiler which is 
properly equipped for that purpose and then to transmit 
it with its contained heat to the engine, which is constructed 
in such a way as to utilize that heat to the best advantage. 
If the steam is to be condensed, as assumed in the ideal 
cases, it has also been found best to remove it from the 

77 



78 



STEAM POWER 



cylinder and to condense it in a separate piece of apparatus 
properly constructed for that purpose. 




The entire arrangement which results from these prac- 
tical modifications in the case of a non-condensing engine 




I 

f///////////////////,...^ 



THE REAL STEAM ENGINE 79 

is shown in Fig. 32. Steam is generated within the boiler 
at some constant pressure P\ and at the proper instant the 
admission valve at one end of the cylinder is opened, allow- 
ing steam to flow in and drive the piston outward. If 
there were no losses, this would be represented by some 
such line as ab at a height Pi on the 
PF-diagram of Fig. 33. Closing of 
the valve after the piston had moved 
part way out would cut off the 
further flow of steam, and, with con- 
tinued motion of the piston, the steam 
within the cylinder would expand. ^ „„ 

If no heat interchanges occurred, 
this expansion be would be adiabatic as in the ideal case. 

It will be observed that the two lines on the PF-diagram 
thus far produced represent equally well the corresponding 
two lines of the complete or incomplete-expansion cycles. 
The heat supplied in the boiler is the same as that supplied 
in the cylinder under the ideal conditions originally assumed, 
and the work under the line a 6 is equal to the external 
work done during vaporization just as in the ideal case. 
If difficulty is experienced in connection with the statement 
regarding external work, it is only necessary to picture the 
process in this way: Assume that each pound of steam 
formed in the boiler does the external work equivalent to 
APu by pushing the pound previously generated ahead of 
it as a piston, and that this motion communicated along 
the pipe from layer to layer results in pushing an equivalent 
weight (and volume) into the cylinder against the resist- 
ance offered to the piston's motion. 

When the piston arrives at the end of its stroke at the 
point c, the opening of the " exhaust valve," connecting 
the interior of the cylinder with the space in which the 
pressure P2 lower than P c is maintained, will permit some 
of the steam to blow out of the cylinder with the piston 
standing stationary at the end of its stroke. This would 



80 STEAM POWER 

give a constant volume change similar to the correspond- 
ing line in the incomplete-expansion cycle. 

The return of the piston from d to e, with the exhaust 
valve still open, would force the remainder of the steam 
out of the cylinder and into the space in which the pres- 
sure P2 is maintained. The result, so far as the diagram 
is concerned, is obviously the same as in the ideal case, and 
if the steam were condensed within a proper vessel into 
which it exhausted (instead of being exhausted to atmos- 
phere), the result would also be the same so far as the shape 
of the diagram is concerned. The pressure P2 might, how- 
ever, be maintained at a lower value, thus giving a greater 
temperature range. 

The pressure rise ea within the cylinder would result 
directly from the opening of the admission valve and the 
admission of steam for the next cycle. But, if the working 
substance is to be returned to starting conditions as was 
dene in the ideal case, its pressure must also be raised to 
Pi and its temperature to a corresponding value. The 
pressure is raised in the case of condensing operation by 
means of the boiler feed pump, which picks up the condensed 
steam (condensate) and forces it into the boiler. The 
temperature of the working substance is raised by passing 
it through feed-water heaters or by heat absorbed directly 
from the heated water in the boiler. 

When operating non-condensing the working substance 
exhausted during the last part of each cycle is really thrown 
away by allowing it to mix with the atmosphere, but an 
equivalent quantity of water is fed to the boiler by the 
boiler feed pump and takes the place of the material lost 
by exhaust to atmosphere. This method of operating 
does not approximate the ideal as closely as does the con- 
densing method, but the discrepancy is not very great. 

53. Losses in Real Installations. The diagram given 
in Fig. 33 was obtained by assuming the absence of certain 
practical losses and is considerably modified when real 



Pt 



THE EEAL STEAM ENGINE 81 

apparatus is used. Thus the real engine, as shown in con- 
nection with Fig. 9, has clearance and operates with com- 
pression so that the clearance is filled with steam at a pres- 
sure indicated by the point a' in Fig. 34 when the admission 
valve opens. 

There is also always some drop of pressure along the 
steam pipe so that the pressure 

at the engine is lower than at the p,| — I — *^*5S!mEv£ 
boiler. Further, the admission g *?\\ /AdiataticExpanBtauoi 

g ^ A>/ all Material in Cylinder 

valve can never be made to give 1 d&q W\ «TnniaiijDryana 

(£ \\y. %V" -^Saturated. 

such a large opening into the \§^j 

cylinder that there is not a P j | Backer, 

measurable drop of pressure in FlG . 34 .-Theoretical and Real 

flowing through it. As a result indicator Diagrams. 

of these actions the highest 

pressure attained within the cylinder as indicated at point 

a in Fig. 34 is always lower than the boiler pressure Pi. 

As the piston of a real engine moves out it acquires 
a higher and higher velocity until it reaches a point near 
mid-stroke. The entering steam therefore must flow 
through the valve with increasing velocity if it is to follow 
up the piston and fill the cylinder, but this usually neces- 
sitates greater pressure drops as the piston moves out, 
so that the admission line generally slopes downward instead 
of being horizontal. There is also another phenomenon 
which causes this line to slope. The metal of cylinder, 
cylinder head and piston is in contact with comparatively 
low-temperature steam during the latter part of each cycle, 
and therefore acquires a lower temperature than that 
of the steam about to enter. Therefore, when the high- 
pressure steam enters the cylinder it gives up heat to 
the walls at a comparatively rapid rate, and, if initially 
dry saturated, this results in a great deal of condensation. 
Such condensation is called initial condensation. 

As the steam condenses after flowing into the cylinder 
and forms water occupying a negligibly small volume, 



82 STEAM POWER 

it follows that steam must flow into the cylinder at a pro- 
portionately greater rate in order to fill the space vacated 
by the piston. But this results in an increased pressure 
drop and therefore would give a sloping admission line. 

When the piston has finally been driven out as far as 
desirable by the action of high-pressure steam, the admis- 
sion valve is closed, that is, cut-off occurs. This valve can 
not be closed suddenly; the closure is more or less gradual 
in all cases. As the opening becomes smaller it becomes 
increasingly more difficult for the steam to flow through 
and into the cylinder so that the pressure continues to drop 
at an increasing rate until the valve is finally closed. This 
gives the rounded cut-off shown at the point b. 

The loss of pressure during admission is generally 
said to be due to throttling or wire drawing, these terms 
being intended to convey the idea that the steam has to 
squeeze its way through the inlet openings with correspond- 
ing loss of pressure. 

When the cut-off has finally been completed, it leaves 
the end of the cylinder filled with a mixture of steam and 
water at steam temperature, and this mixture then expands 
as shown by the line be. At the beginning of the expansion 
the steam generally has a higher temperature than that of 
the surrounding walls and it therefore continues to give 
heat to those walls. Were the expansion adiabatic it 
would follow the dot-dash line in the figure, but, as the 
steam must not only convert heat into work, but must also 
supply heat to the walls, it condenses more rapidly than 
in the ideal case and its pressure and volume changes follow 
some such law as that indicated by the upper part of the 
curve be. 

As expansion continues, the pressure and temperature 
of the steam drop until some point is reached at which 
the temperature has become equal to that of the walls. 
Further expansion with drop of pressure and temperature 
results in reducing the temperature of the steam below 



THE REAL STEAM ENGINE 83 

that of the walls, and then the direction of heat transfer 
is reversed, the hot walls giving heat to the cooler steam 
at an increasingly rapid rate. This heat causes re-evapora- 
tion of some of the water formed before and thus tends to 
increase the volume occupied by the material in the cylinder, 
with the result that the lower part of the expansion curve 
be approaches and generally crosses the curve which would 
have been attained by adiabatic expansion in non-conduct- 
ing apparatus. 

In many real engines the re-evaporation is so great that 
the steam is entirely dried and sometimes superheated 
before the exhaust valve opens. 

The exhaust valves of steam engines are always opened 
before the piston reaches the end of the stroke, as it is found 
necessary to do this if excessive losses are not to occur 
due to the difficulty of forcing the large volume of low- 
pressure steam through the exhaust passages. When opened 
early enough, the steam flows out in such quantity before 
the end of the stroke that the " back pressure " during the 
return or exhaust stroke is only a pound or two above that 
of the space into which the engine is exhausting. 

During all of the exhaust period, the steam is probably 
at a lower temperature than the walls to which it is exposed 
and re-evaporation probably continues in most cases until 
the closure of the exhaust valve. It seems probable that 
the steam retained in the cylinder after the closure of the 
exhaust valve is approximately dry, but little is really 
known regarding the quality of the clearance steam. 

The rise of pressure during compression has two bene- 
ficial effects : It helps to bring the moving parts to rest grad- 
ually, and it raises the temperature of the clearance steam 
and of the walls of the clearance space to values nearer 
that of the entering steam. 

Remembering that area on a PF-diagram represents 
work, it is easily seen that throttling losses and rounding 
of corners due to slow valve action (which cause a loss of 



84 STEAM POWER 

diagram area) result in a loss of work. The fact that conden- 
sation also causes a great loss is easily shown. A given 
quantity of steam entering the engine with its supply of 
heat can, in the ideal case, do a certain amount of work at 
the expense of that heat. In the real case part of the heat 
is stored in the walls during the early part of the cycle, 
so that it is not available for the doing of work and is removed 
from the walls and carried out into the exhaust as unutilized 
heat during the later part of the cycle. The phenomenon 
can be pictured by imagining the steam as dropping some 
of its heat into a pocket in the walls of the cylinder when 
entering the engine and then picking it up again and carrying 
it out when leaving, so that the next charge of steam will 
have to fill the pocket again. 

The net result of condensation and re-evaporation 
is the obtaining of less work from a given quantity of steam 
than should be obtained, or the use of more steam than 
theoretically necessary for a given quantity of work. This 
effect is shown graphically by the two adiabatic expansion 
lines of Fig. 34. 

The initial condensation in real engines which are sup- 
plied with saturated steam generally amounts to from 20 to 
50 per cent of all the steam supplied, so that it is evident 
that anything which will prevent part or all of this loss should 
do much to improve the steam consumption of engines. 
This subject will be discussed in more detail in later para- 
graphs and various methods of decreasing losses from this 
source will be considered. 

54. Clearance. The term clearance is used in a two- 
fold sense; (a) to refer to mechanical clearance or the 
linear distance between the two nearest points of cylinder 
head and piston face when the piston is at the end of its 
stroke, and (6) to refer to volumetric clearance or the 
volume enclosed between the face of the valve, the cylinder 
head and the face of the piston when the latter is at the 
end of its stroke. 



THE REAL STEAM ENGINE 



85 



Steam Port Steam Chest Valve 




Fiston 



Fig. 35. — Mechanical and Volu- 
metric Clearances. 



The former is generally given in inches and varies 
from a very small fraction of an inch in the best engines 
to an inch or more in cheap and in poorly designed engines. 
It is indicated by a in Fig. 35. 

The volumetric clearance is expressed as a percentage 
of the piston displacement or 
volume swept through by the 
piston. It varies from 2 per 
cent or less in the best engines 
to as high as 15 per cent in 
the cheaper and less economical 
models. It is made up of the 
parts designated by c in Fig. 35. 

55. Cushion Steam and Cyl- 
inder Feed. It is customary 
to imagine the steam operating 
within an engine cylinder to 
consist of two parts, the cushion 
steam and the cylinder feed. The former is that part of 
the total which is contained in the clearance space before 
the admission valve opens and serves to cushion the 
reciprocating parts of the engine. The cylinder feed 
is the steam which enters through the valve for each 
cycle. 

If the same cycle is produced time after time so that all 
temperature effects are repeated at regular intervals and 
so that all events occur at the same points in successive 
cycles, the quantity of steam retained in the clearance 
volume will be the same for successive cycles. It is 
impossible to measure the quantity of this steam directly 
and indirect methods are therefore adopted for that 
purpose. 

It is often assumed that the steam is dry and satu- 
rated when compression begins, as at the point e in Fig. 
34. With this assumption, the weight of cushion steam 
can be determined by dividing the volume occupied, that is, 



86 STEAM POWER 

Ve t by .the volume occupied by one pound of dry saturated 
steam at the same pressure. Thus, 

Cushion steam = = ; — - — =- lbs. . (29) 

Sp.vol. at pressure P e 

The weight of cylinder feed can be very accurately 
determined by condensing and weighing the steam leaving 
the engine in a given time and dividing by the number of 
cycles performed during the same period. It can also be 
determined by metering the steam entering the engine 
or by measuring the water fed to a boiler supplying only the 
engine in question. An approximate determination of the 
quantity of the cylinder feed can also be made directly from 
an indicator diagram by determining what is known as the 
diagram water rate. This will be considered in detail at 
a later point. 

When cushion-steam and cylinder-feed have both been 
determined, the weight of steam contained in the cylinder 
between cut-off and release can be found by adding the two 
quantities. Thus, 

W = W f +W K , (30) 

in which 

W — total weight of steam expanding in cylinder per cycle; 
Wf= weight of cylinder feed per cycle; and 
Wk = weight of cushion steam per cycle. 

The volume which the mixture would occupy if dry 
and saturated at any given pressure can be determined by 
multiplying W the total weight by the specific volume for 
that particular pressure. 

56. Determination of Initial Condensation. The loss 
due to initial condensation is so important that it is cus- 
tomary to determine the amount of this loss when studying 
engines. This can be done with fair accuracy by means 
of the indicator diagram. 



THE REAL STEAM ENGINE 



87 



To make such a study it is necessary to know the total 
freight of material in the engine cylinder at the point of 
cut-off. This weight may be determined by any of the 
methods just given. With the weight known, the volumes 
which this material should occupy at different pressures 
if dry and saturated can be determined by multiplying by 
the specific volumes at the various pressures. Plotting 
these points on a PF-diagram and connecting them will 
give a saturation curve for the material in the cylinder such 
as the curve shown in Fig. 36. 

By drawing this curve on the indicator diagram ob- 
tained from the engine and then comparing distances 
such as ab and ac as explained in section 26 of Chapter III 





Fig. 36. 



Fig. 37. 



the quality of the steam within the cylinder at all pressures 
between cut-off and release can be determined. The weight 
of initial condensation (up to the point of cut-off) must 
be the total weight of water shown as existing within the 
cylinder at that point minus any water brought in by the 
steam if it was not dry when entering the engine. 

Should the saturation curve cross the real expansion 
curve, as shown in Fig. 37, it indicates that the steam oc- 
cupies volumes greater than the specific volumes toward 
the end of the expansion; the steam within the cylinder 
must therefore be superheated during this part of the 
cycle. 

Many formulas have been devised for giving the quan- 
tity of initial condensation. They are all based upon the 
results of experiment and generally only give reliable 



88 STEAM POWER 

values for cases similar to those used in developing them. 
One formula of this sort which has been very widely tested 
and been found to give reliable results within its field 
of applicability is that devised by Robert C. H. Heck and 
explained in his books on the steam engine. The formula is 



c_ sd 
\fN\pe 
in which 



=«fe 



ra = the fraction representing initial condensation; for 
ordinary cases it is the fraction of the cylinder 
feed which is condensed during admission, but wheu 
compression is very high and when great weights 
of steam are retaiued in the clearance it is the frac- 
tion of all the material within the cylinder which 
exists in liquid form at the time of cut-off; 
c = a coefficient, which varies between 0.25 and 0.30 with 
ordinary engine types. Its value is unknown for 
certain new types such as the Una-flow. 

N = engine speed in revolutions per minute (R.P.M.); 
s = a constant for any engine, equal to nominal surface in 
square feet divided by nominal volume in cubic 
feet. The nominal surface is the area of the inner 
walls and the ends of a cylinder with diameter equal 
to the internal diameter of the cylinder and with 
a length equal to the stroke of the engine. The 
nominal volume is the cubic contents of such a 
cylinder; 

s = y-(2-^-f-4j in which D and S represent diameter and 

stroke of engine in inches; 
= a temperature function obtained from Table II as 

there indicated; 
p = the absolute pressure in cylinder in pounds per square 

inch just after completion of cut-off; 



THE REAL STEAM ENGINE 



89 



e = cut-off ratio, that is, ratio of cylinder volume opened 
up by time cut-off has just been completed, to the 
total piston displacement. 

TABLE II 

For Finding Values of .0 for Use in Heck Formula 

ki~ kz when ki and ki are chosen from table for highest and lowest pressures 
existing in cylinder 



V 


k 


V 


k 


V 


k 


p 


k 


V 


k 


V 


k 


1 


175 


15 


210 


50 


269.5 


90 


321.5 


160 


389 


230 


441 


2 


179 


20 


220 


55 


277 


100 


332.5 


170 


397 


240 


447.5 


3 


183 


25 


229 


60 


284 


110 


343 


180 


405 


250 


454 


4 


186 


30 


238 


65 


291 


120 


353 


190 


413 


260 


460.5 


6 


191 


35 


246 


70 


297.5 


130 


362.5 


200 


420 


270 


467 


8 


196 


40 


254 


75 


304 


140 


371.5 


210 


427 


280 


473 


10 


200 


45 


262 


80 


310 


150 


380.5 


220 


434 


290 


479 



57. Methods of Decreasing Cylinder Condensation. 

Before discussing methods of decreasing the loss due to cylin- 
der condensation it will be well to consider what things may 
be expected to determine the extent of such loss. The 
condensation is due directly to the transfer of heat from 
one body to another at lower temperature, and anything 
which tends to increase the total amount of heat thus trans- 
ferred will increase the total condensation. 

It is therefore evident that the ratio of steam condensed 
to steam supplied will be greatest when: 

(a) The time of contact is greatest; 

(b) The ratio of surface exposed to volume enclosed is 
greatest, and 

(c) The temperature difference is greatest. 

The time of contact can be controlled to a certain 
extent by controlling the speed of the engine and, with 
other things equal, the higher the speed the lower should 
be the condensation. 

The ratio of surface exposed to steam to the volume 
occupied by steam has a great influence on the amount of 



90 STEAM POWEK 

condensation which occurs. The surface of the clearance 
space, including the interior surfaces of all ports or passages 
leading to the valves, seems to have the greatest influence, 
and the clearance space which is most nearly a short cylinder 
without connected passages may be expected to give the 
least initial condensation. 

The size of the engine is also important in this connec- 
tion. The area exposed does not increase as rapidly as 
does the volume inclosed when the diameter of a cylinder is 
increased, and therefore large cylinders give smaller ratio 
of surface to volume and therefore a smaller percentage 
of steam condensed. Large engines thus have a decided 
advantage over small engines. 

The shape of the cylinder also has an effect. The 
longer the cylinder with respect to its diameter the more 
favorable its performance. 

The point at which cut-off occurs is also intimately 
connected with the condensation loss. In a given cylinder 
with a given clearance the total condensation within the 
clearance space may be assumed practically constant if 
speed and temperature remain about the same. But if 
the cut-off is made later larger quantities of steam are 
admitted per stroke, and hence the fraction of the total 
cylinder feed which is condensed decreases. 

The temperature differences depend on upper and 
lower pressures, that is, on the pressure range. The inner 
surfaces of the walls follow as rapidly as possible the tem- 
perature changes of the steam within them. Thus their 
average temperature is somewhere between the upper and 
lower temperatures of the steam. If now, with a given 
upper steam pressure and therefore temperature, the lower 
pressure be reduced, the average wall temperature also will 
be reduced, and therefore the differences between the 
temperature of the entering steam and the average tem- 
perature of the walls will be increased with a resulting in- 
crease in condensation loss. 



THE REAL STEAM ENGINE 91 

The methods of decreasing this loss can now be con- 
sidered. They are given below under separate heads 
with brief explanation when necessary. 

(a) Clearance spaces should be properly designed so 
that the minimum surface is exposed. 

(6) The proportions of cylinder (diameter and stroke) 
and the speed of the engine should be so chosen that the 
condensation loss is reduced to a minimum. 

(c) The engine should be so proportioned that when 
delivering its rated power the cut-off occurs at such a point 
as to make the percentage of cylinder condensation the 
minimum consistent with other requirements. 

(d) The cylinder should be surrounded by spaces filled 
with air or by materials which are poor conductors of heat 
so as to decrease loss by radiation, because all heat lost in this 
way must be supplied by the condensation of steam within 
the cylinder. Such metallic parts as cannot be " lagged" 
in this way should be polished because polished surfaces 
radiate less heat than dull surfaces under like conditions. 

(e) The cylinder may be surrounded by a steam jacket, 
that is, a space filled with steam similar to that supplied 
the cylinder. The use of such a jacket sometimes results 
in a considerable saving and at other times in a great loss. 
The cylinder proportions, speed and pressure range seem 
to be the determining factors, and most long-stroke cylin- 
ders operating at low rotative speed and with small pressure 
ranges are jacketed. 

(/) The engine may be compounded, that is, the expan- 
sion of the steam may be made to occur in two or more 
cylinders taking steam in series. This results in decreas- 
ing the pressure range in each of the cylinders and effects 
a decided saving under proper conditions. Compounding 
will be considered in detail in a later chapter. 

(g) The engine may be supplied with superheated steam. 
If the steam is sufficiently superheated it can give up part 
or all of its superheat to heat the cylinder walls, and thus no 



92 STEAM POWER 

condensation need occur. Heat interchanges between 
metal and superheated steam also appear to be less rapid 
than is the case when the steam contains water, so that a 
saving results from this source also. 

Tests made with saturated and with superheated steam 
indicate that from 7° to 10° of superheat are generally 
required to prevent 1 per cent of initial condensation. 
Results differ greatly with the character of the engine, with 
its economy on saturated steam, with its valve gear, etc. 
Superheats of from 25° to 50° can generally be used with 
any well-designed engine, but higher temperatures usually 
require specially constructed engines. With proper cylinder 
and valve construction the maximum permissible temperature 
is set by the lubricating oil. At the present time the max- 
imum steam temperature is thus limited to 600° F. or less. 

(h) The engine may be built to operate on the Una-flow 
cycle described later. In this case a very ingenious modifica- 
tion of the engine results in decreasing the temperature dif- 
ferences between walls and steam, with resulting diminution 
of heat transfer. 

58. Classification of Steam Engines. Steam engines, 
are classified on many different systems, the one used in 
any particular case being determined largely by circum- 
stances. The principal methods of classification are indi- 
cated in the following schedule : 

Classification of Steam Engines 

On the basis of rotative speed. 

(a) Low speed; (6) Medium speed; (c) High speed. 

On the basis of ratio of stroke to diameter. 
(a) Long stroke; - (6) Short stroke. 

On the basis of valve gear. 

A. Slide valve; (a) D-slide valve; (b) Balanced slide 
valve; (c) Multiported slide valve; (d) Piston 
valve. 



THE REAL STEAM ENGINE 93 

B. Corliss valve; (a) Drop cut-off; (b) Positively 

operated. 

C. Poppet valve. 

On the basis of position of longitudinal axis. 

(a) Vertical; (6) Inclined; (c) Horizontal. 

On the basis of number of cylinders in which steam expands. 

A. Single expansion or simple engine. 

B. Multi-expansion engine, (a) Compound expansion; 

(b) Triple expansion ; (c) Quadruple expansion. 

On the basis of cylinder arrangement. 

(a) Single cylinder; (6) Tandem compound; (c) 
Cross compound; (d) Duplex. 

On the basis of use. 

(a) Stationary engines; (b) Portable engines; (c) 
Locomotive engines; (d) Marine engines; (e) 
Hoisting engines. 

59. Rotative Speeds and Piston Speeds. High-speed 
engines operate at a comparatively high rotative speed and 
are characterized by a short stroke in comparison with the 
diameter of the cylinder, the stroke generally being equal 
to, or less than, the diameter. The piston speed, by which 
is meant the feet travelled by the piston per minute, generally 
falls between 500 and 700. 

It is not considered advisible to allow piston speeds of 
stationary steam engines to exceed about 750 feet per 
minute for ordinary constructions and the great majority 
of engines give much lower values. The piston speed will 
obviously be given by the formula 

S = 2LN, (32) 

in which 

S = piston speed in feet per minute; 
L = stroke in feet; and 
N = revolutions per minute, 



94 



STEAM POWEK 



and it is evident from this formula that as the rotative 
speed is increased the piston speed will increase unless 
the length of stroke is proportionately decreased. As 
a result, high-speed engines have short strokes in com- 



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375 
350 
325 
300 

275 

250 

225 * 

200 

175 

150 

125 

100 



8 10 12 14 16 18 20 22 24 26 
Diameter of Cylinder (Inches) 

Fig. 38. — Proportions of High Speed Engines. 



parison with their cylinder diameters and slow-speed 
engines have long strokes. 

The characteristic relations between cylinder diameter 
and stroke, rotative speed and piston speeds of high-speed 
engines are given in Fig. 38. 

High-speed engines are generally fitted with some 



THE REAL STEAM ENGINE 



95 



"S 8 



R. P. M. 



Piston Speed (S) Ft. per Min. 

S o »o o >ra o 

so CO W3 »o -* -h 



















































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a o 

CM 

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3 3 8 
Stroke (L) in Inches 



96 STEAM POWER 

form of balanced slide valve, and are controlled by what 
is called a shaft governor. They are very compact, having 
small weight and occupying small space in comparison with 
the power developed. 

Slow-speed engines are, in general, the most economical 
and are characterized by low rotative speed, long stroke, 
and elaborate valve gear. The weight per horse-power 
is high, and they generally occupy a great deal of space. 
The Corliss engine is the best known and most widely built 
engine of this type in this country. 

The characteristic relations of cylinder diameter to 
stroke, rotative speed and piston speed for slow-speed 
engines are given in Fig. 39. 

Medium-speed engines generally operate at rotative 
speeds between 150 and 250 R.P.M. They are generally 
fitted with the better forms of multiported and balanced 
slide valves, with poppet valves, or with a positively operated 
Corliss type of valve. 

60. The Simple D-slide Valve Engine. The simplest 
and cheapest type of reciprocating steam engine manu- 
factured is shown in part section in Fig. 40 with the prin- 
cipal parts labelled. The cylinder, piston, steam chest 
and valve are sectioned in order to show the internal con- 
struction. 

This engine, like most steam engines, is double acting, 
that is, a cycle is produced on each side of the piston during 
every revolution. Steam is admitted and expanded on one 
side of the piston while steam is being exhausted on the other 
side. The control of admission and exhaust is effected by 
the slide valve and will be considered in detail in later sec- 
tions. 

The mechanical energy made available by the steam 
operating in an engine cylinder is not developed at a uniform 
or constant rate, but fluctuates, during each revolution, 
above and below the amount required to overcome the 
constant resistance at the shaft due to the work the engine 



THE REAL STEAM ENGINE 97 

is doing. If no provision were made to prevent it, this 
wo aid result in a very variable rate of relation during 
each revolution. When the energy made available was in 
excess of the demand it would be used in accelerating 
the moving parts of the engine and the speed of the latter 
would increase. The reverse would occur when the supply 
did not equal the demand. 

The fly-wheel is used to prevent violent fluctuations 



Cornbined Flywheel 
and Beltwheel 




Fig. 40. — Simple D-slide Valve Engine. 

of this kind. It is made with a comparatively heavy rim 
and a great deal of energy must be supplied to accelerate 
it to any appreciable extent in a short time. Similarly 
it can give out a great deal of energy when slowing down. 
The fly-wheel therefore serves as a sort of reservoir in which 
excess energy can be stored temporarily and from which 
it can later be withdrawn when a deficiency exists. The 
fly-wheel thus acts as a damper to variation of rotative 



98 



STEAM POWER 



speed during each revolution, minimizing but not entirely 
eliminating such variation. It may also serve as a belt 
wheel, as shown in the illustration. 

The governor controls the steam supply to the cylinder 
in such a way that enough heat will be supplied to make 
available the power demanded at the shaft. Were more 
supplied the excess would be absorbed by the moving 
parts and the engine speed would increase, were less supplied 
the engine speed would decrease. 

61. Engine Nomenclature. The meanings of several 
terms used in describing engines are not self-evident, their 



Belt Backward 



Belt Forward 



'Left Hand 
Engine 




Fig. 41. — Engine Nomenclature. 



definitions depending merely on accepted usage. Some 
of these terms and their meanings are illustrated in Fig. 41. 

The crank end of a horizontal engine is called the front 
of the engine, so that the cylinder head nearest the crank 
is called the front head and the stroke of the piston toward 
the crank is known as the forward stroke. The forward 
stroke of the piston is also spoken of as the outstroke, par- 
ticularly in connection with single acting engines. The 
stroke away from the crank is correspondingly designated 
as the return or the instroke. 

62. Principal Parts of Engines. The parts of engines 
may be roughly divided into stationary and moving, such as 
frame, cylinder, cylinder and valve chest covers, etc., which 



THE EEAL STEAM ENGINE 



99 



are stationary, and piston, piston rod, crosshead, fly-wheel, 
etc., which are all moving parts when the engine is in opera- 
tion. The moving parts are often divided into reciprocal- 




Fig. 42. — Frame for Small Vertical Engine. 

ing and rotating parts. Thus the piston and all connected 
parts through and including the crosshead, and the valve 
and many connected parts in the case of slide valve engines, 



Jaws for 
main bearings 




Fig. 43. — Frame for Medium Speed Center Crank Engine. 



all reciprocate when the engine is in operation. The 
shaft, fly-wheel, eccentric sheaves and governor constitute 
the principal rotating parts. 



100 



STEAM POWER 



Some engines also have oscillating parts, such as the 
valves in Corliss types, which rock back and forth in the 
arc of a circle, and the rocker arms in various forms of 
valve gear, these arms rocking through a short arc about 
a fixed pin wear one end. 

The principal parts and their functions are briefly con- 
sidered in the following paragraphs: 

(a) The Frame. The frame of the engine, sometimes 
known as the bed, serves to support the other parts, to tie 



Cylinder bolted 
to finished 
surface here. 




Jaws for 
main bearing 



Fig. 44. — Frame for Slow Speed Engine of Corliss Type; 
Overhung Crank. . 



Side or 



them together in their proper relations and to fasten the 
whole structure to whatever foundation is used. The cross- 
head guides and the seats for the main bearings are incor- 
porated in the frame. 

The frame is commonly made of cast iron in the form 
of a hollow box which is properly ribbed to give the neces- 
sary stiffness. 

Examples of frames are shown in Figs. 42, 43 and 44. 



THE REAL STEAM ENGINE 



101 



(6) The Cylinder and Steam Chest. The cylinder and 
steam chest are generally incorporated in the same casting, 
and surfaces of covered cavities in this casting are finished 



^SteamPlpe 




Steam Cliest 
Cover 



Fig. 45. 

-Steam Chest Cover 



Steam 
Chest 




Cylinder 

Fig. 46. 



to form the cylinder in which the piston operates and the 
seat or seats upon which valves rest and move. 

The cylinder may be single walled with flanges on the 
end to receive the cylinder head, as illustrated in Fig. 40 



102 



STEAM POWER 



(plain D-slide valve), in which case a thin sheet-metal 
jacket is fastened around it and the space between filled 
with heat-insulating material. Or, the cylinder may be 
cast with double walls, the space between the two being 
used as an air jacket or as a steam jacket. 

Some cylinders are fitted with a liner, which is a plain 
cylinder pressed into place within the cylinder casting 
and forming the bore of the working cylinder. This prac- 




Fig. 47. — Section of Atlas Medium Speed Engine, Showing Balanced 

Slide Valve. 

tice is common on the larger types, the liner being used so 
that when wear has occurred it can be replaced cheaply, 
instead of it being necessary to rebore or even replace 
the main casting. 

Examples of cylinder construction are shown in Figs. 
45, 46, 47, 48, 49 and 50. 

(c) The Piston. The function of the piston is two- 
fold. It must prevent the leakage of steam by it from one 
end of the cylinder to the other, and it must receive the 



THE REAL STEAM ENGINE 



103 




-3 
M 



104 



STEAM POWER 



pressures exerted by the steam and transmit them to the 
other parts of the mechanism as it moves. 



Chambers for 

Steam Valves 

(admission valves) 



Chambers for 
Exhaust Valves 




Steam Port 



Surface of 
Cylinder Bore 



Exhaust Port 



Fig. 49. — Corliss Cylinder Casting. 

-Steam Connection 




Fig. 50. — Corliss Cylinder with Lagging in Place. 



Leakage of steam is prevented by the use of piston 
rings, which are metal rings fitted into grooves in the circular 
surface of the piston and pressed out against the cylinder 




THE REAL STEAM ENGINE 105 

Pistons 




Fig. 51. 



(a) (6) 

Fig. 52. 




End of 
Piston Rod' 



■^BuH Rings 

Fig. 53. — Built-up Piston Used in Large Engines. 



106 



STEAM POWER 



walls by spring action. They may be made of one piece 
of metal turned into a ring of slightly larger diameter than 
the cylinder, cut through and sprung into place, or they may 
be made in pieces as shown in Fig. 51, and pressed out 
against the wall by small helical or leaf springs. 

The piston itself may consist of a solid disk of metal 
fitted with a hub and a short cylindrical part with grooves 
for the rings, as shown in Fig. 52 (a) and (6), or it may be 
an elaborate built-up structure, as shown in Fig. 53. 




Fig. 54. — Crosshead and Pin. 



(d) The Piston Rod. The piston rod is a plain cylindrical 
steel rod fitted with such shoulders and threads at the ends 
as are necessary for the fastening of the piston and the cross- 
head respectively. Examples of such fastenings are given 
in Figs. 51, 52, 53 and 55. 

In large horizontal engines the piston rod sometimes 
extends through the piston and rear cylinder head, and 
the rear end is then supported by a small auxiliary cross- 
head. The extension of the rod is known as the tail rod 
and the auxiliary crosshead as the tail rod crosshead. 



THE REAL STEAM ENGINE 



107 



Such constructions are used when the weight of the piston 
is so great that it would cause serious cylinder wear if not 
supported more perfectly than is possible with the ordinary 
overhung arrangement. 





Fig, 55. — -Single Slipper Crosshead. 




Fig. 56. — Crosshead with Adjustable Slippers. 

(e) The Crosshead and Guides. The crosshead and 
guides are used for the purpose of supporting the piston 
and its rod and guiding them in a straight line. The 
crosshead also serves to connect the piston rod and the 
connecting rod through which the forces are transmitted 
to the crank pin. 



108 



STEAM POWEE 



Crossheads are generally cast in the form of imperfectly 
shaped boxes and carry slippers which are faced with anti- 
friction metal where they come in contact with the guides. 
The slippers may be flat and operate on planed guides, 
as shown in Fig. 40, or they may be turned and operate in 
bored guides, as shown in Figs. 48, 54, 55 and 56. Pro- 
vision is generally, though not always, made for taking 
up wear of guides and slippers by setting the slippers 
further out from the body of the casting. 

With the type shown in Figs. 40 and 54 this can be 




Fig. 57. — Solid End Connecting Rod; for Overhung Cranks only. 



done by the insertion of thin sheets of metal or paper 
(known as shims) between the body of the crosshead and the 
slippers. In the type shown in Fig. 56 the slippers are 
finished with inclined surfaces where they come in contact 
with the main casting, and the adjustment is made by wedg- 
ing the slippers apart by the use of the adjusting bolts 
shown. 

The wrist-pin end of the connecting rod enters the 
crosshead casting and is held in place by means of the 
wrist pin, about which it oscillates when the engine is in 
operation. 



THE REAL STEAM ENGINE 



109 



(/) The Connecting Rod. This rod connects the re- 
ciprocating crosshead with the rotating crank shaft and 
transmits the forces from one to the other. It consists 
of a body or shank and two ends or heads. The ends may 




lUILUJ 



Fig. 58. — Connecting Rod with Bolted Strap Ends; May be Used 
with Center or Side Crank Constructions. 

be " closed " or " solid " as shown in Fig. 57; they may- 
be made with a strap bolted in place as shown in Fig. 58; 
or the crank-pin end may be made of two half boxes bolted 
together to form a " marine end " as shown in Fig. 59. 
The ends are always made adjustable so that wear 



Crank End 



Crosshead End. 




Fig. 59. — Marine End Connecting Rod. 



can be taken up, thus preventing noisy operation due to 
hammering between the ends and the pins at times when 
the direction in which forces act is reversed. With solid 
and strap types this adjustment is generally made by 
means of wedges similar to those shown in Figs. 57 and 58. 



110 



STEAM POWER 



With the marine type shims are used between the two 
halves, and the diameter of the hole formed by the latter 
is decreased by the removal of shims of the required thick- 
ness. 

(g) The Shaft. The crank shaft itself is generally made 



Cran'kPin 

Crank. Arms 




"Journals' 

Fig 60.— Crank Shaft, Center Crank. 




Counterweights 
or Balances 

Fig. 61. — Center Crank. 




Counterweight 

Fig. 62.— Center Crank. 



of steel, but the counterbalances are often of cast iron. It 
may be one forging throughout or may be built up by 
shrinking the various parts together. Multicrank shafts 
of large size are generally of built-up construction, the crank 
pins being shrunk into the crank arms and the latter shrunk 
on to the pieces of shaft. 

The counterbalance weights are used to balance the 



THE REAL STEAM ENGINE 



111 



centrifugal effect of the crank pin, part of the crank arms 
and part of the connecting rod, all of which rotate off center. 
In some engines part of the unbalanced effect of the recip- 
rocating parts is also imperfectly balanced by these counter- 
balances. 

Various types of shafts are shown in Figs. 60, 61, 62, 
63, and 64. 

(h) Bearings. Bearings are distinguished as main and 
as outboard bearings. Main bearings are those carried by 




Fig. 63. — Crank Shaft and Disc, Overhung Crank. 




Fig. 64. — Overhung Crank. 



the frame of the engine and outboard bearings are carried 
by separate pedestals or by pedestals fastened to a plate 
which is in turn fastened to the frame. Center-crank 
engines have two main bearings, and side- crank engines 
have only one, the other end of the shaft being supported 
by an outboard bearing. 

The bearings of steam engines are generally formed 
of babbitt-lined boxes carried within jaws machined in a 
frame, or in a separate pedestal, and held in place by a 
bearing cap. The boxes are made in two, three or four 
parts to allow for adjustment to compensate for wear and 



112 



STEAM POWER 



to give a certain degree of flexibility. Adjustment for 
wear is either made by means of wedges or by means of 
screws which force the various parts of the boxes toward 
the shaft. An example of a three-part bearing with screw 
adjustment as used with large side-crank engines is shown 
in Fig. 65. The parts of a three-part bearing with wedge 
adjustment are shown in Fig. 66. 

Bearings are often lubricated by rings or chains, and 
they are then known as ring- or as chain-oiling bearings. 




Fig. 65. — Three Part Main Bearing with Screw Adjustment. 



In the ring-oiling bearing one or more metal rings of large 
diameter hang loosely on the shaft within the bearing and 
dip into an oil reservoir below the shaft. Rotation of the 
shaft causes the rings hanging on it to rotate and they 
carry oil up from the reservoir and spill it out over the shaft 
within the bearing. Chain bearings are similar except that 
chains are substituted for rings. 

(i) Fly-wheels. The function of the fly-wheel has al- 
ready been considered and need not be discussed further. 
The wheel is constructed with a heavy rim joined to a hub 



THE REAL STEAM ENGINE 



113 



by six or eight arms. In the smaller sizes the wheel may be 
cast in one piece, but best practice calls for a split hub in 
that case to partly equalize certain casting strains which 
result from unequal thicknesses of metal in different parts 
of the wheel. Large wheels are cast in two or more parts 
both for the purpose of partly avoiding casting strains 
and for the purpose of facilitating handling and shipping. 





Fig. 66. — Three Part Bearing 
Showing Wedge Adjustment. 



Fig. 67. 



A two-part wheel with the rim sections joined by 
prisoner links shrunk in place and the hub fastened with 
bolts is shown in Fig. 67. 



PROBLEMS 

1. A given engine has a piston displacement of 3 cu.ft. and a 
clearance volume of 3%. Compression begins when 85% of the 
exhaust stroke has been completed and the pressure within the 
cylinder at that time is 16 lbs. per square inch absolute. Deter- 
mine the weight of the cushion steam on the assumption that this 
steam is dry and saturated at the beginning of compression. 



114 STEAM POWER 

2. Assume the engine described in Prob. 1 to cut-off at J stroke 
and with a pressure inside the cylinder equal to 115 lbs. per square 
inch absolute. Find the weight of ^cylinder feed if the quality 
of the material in the cylinder at the'time of cut-off is 75%. 

3. Find the piston speed of an engine with a stroke of 2 ft. 
and a rotative speed of 150 R.P.M. 

4. Show by means of Heck's formula that initial condensation 
increases with pressure range. 



CHAPTER VIII 
THE INDICATOR DIAGRAM AND DERIVED VALUES 

63. The Indicator. The ideal steam engine cycle was 
described in Chapter IV, and the sort of diagram which 
would be obtained from a real engine was shown in Chapter 



Drum 



-Point Holder 




Cylinder — > 



Fig. 68. 



VII; but the means by which such diagrams are obtained 
from operating engines was not given. 

Indicator diagrams showing the pressure and volume 
changes experienced by steam in the cylinders of real 

115 



116 



STEAM POWER 



engines are obtained by means of an instrument known as 
an indicator. The operation of obtaining such diagrams is 
known as indicating the" engine. 

An external view of one form of indicator is shown in 
Fig. 68 and a section through the instrument is given in 
Fig. 69. The method of connecting an indicator to the 



Pi'ston Rod- 




"Connected to Engine 
Cylinder 

Fig, 69. 



cylinder of a steam engine and one method used for driving 
it are illustrated in Fig. 70. 

The indicator is intended to draw a diagram showing 
corresponding pressures and volumes within the engine 
cylinder and must, therefore, contain one part which will 
move in proportion to pressure variations and another which 
will move in proportion to volume changes. The one may 



INDICATOR DIAGRAM AND DERIVED VALUES 117 

be called the pressure-measuring and the other the volume- 
measuring device. 

The pressure-measuring device generally consists of 
a piston, such as shown in the figure, working with minimum 
friction in a small cylinder and fitted with a spring which 
will resist what may be called outward motion (upward 
in the figure). The cjdinder containing this piston is 
coupled to a short pipe connected with the clearance space 
of the engine and, whenever the indicator cock in this 
connection is open, the steam acting on the engine piston 




Fig. 70. — Method of Attaching and Operating an Indicator. 

will also act on the indicator piston. Steam of any given 
pressure will drive the indicator piston out against the 
action of the spring until the pressure exerted by the spring 
is equal to that exerted on the face of the piston. The 
indicator piston will thus move out different distances for 
different pressures, and, through the piston rod and pencil 
mechanism, will move the pencil point to various heights 
corresponding to different steam pressures. The pencil 
mechanism is so arranged that the point traces a straight 
vertical line on the drum as the indicator piston moves in 
and out. 

Springs are made to certain definite scales, thus there 



118 STEAM POWER 

are, for instance, 10-lb., 25-lb., 50-lb. and 100-lb. springs. 
The number which is known as the scale of the spring 
designates the steam pressure in pounds per square inch 
which is required to move the pencil point 1 inch against 
the action of such a spring. With a 100-lb. spring in the 
indicator, a steam pressure of 50 pounds per square inch 
acting on the indicator piston would drive the pencil up a 
distance of half an inch, a pressure of 100 pounds per square 
inch would give 1 inch of motion and so forth. 

The volume-measuring device is of an inferential kind. 
It simply indicates the position attained by the engine 
piston at the time when a given steam pressure existed in 
the cylinder and the volume occupied by the steam can be 
calculated from piston position and cylinder dimensions. 
The position of the piston is indicated by connecting the 
cord wound around the drum to some part of the engine 
which is rigidly connected to the piston. The crosshead 
is commonly used for this purpose and, since the motion 
of this member is generally much greater than the circum- 
ference of the drum, it is necessary to use a reducing 
mechanism of some sort. This mechanism must be very 
accurate, so that it moves the drum as nearly as possible 
in proportion to the motion of the engine piston. 

The pencil point moves up and down as the pressure 
within the cylinder varies, and the drum rotates under 
the point in proportion to the motion of the engine pis- 
ton, so that the combination of the two motions brings the 
pencil point to successive positions on the drum which indi- 
cate successive corresponding values of steam pressure and 
piston position. By mounting a piece of paper, known 
as a card, on the drum and pressing the pencil point upon 
this paper, the successive positions occupied by the pencil 
point will be recorded in the form of a series of curves and 
straight lines. 

If the drum is rotated with the lower side of the indica- 
tor piston connected to atmosphere, the pencil will trace 



INDICATOR DIAGEAM AND DERIVED VALUES 119 

a horizontal line. This is known as the atmospheric line 
and is used as a reference for locating the pressure scale. 
If the indicator cylinder is then connected with the engine 
cylinder and the drum is rotated by the reducing mechanism, 
a diagram similar to that of Fig. 71 will be drawn upon the 
card. The atmospheric line indicates the height assumed 
by the pencil when atmospheric pressure acts on the 
piston and, knowing the value of the existing atmospheric 
pressure (barometer reading) and the scale of the spring, 
a line at a height representing zero pressure can be drawn 
on the card. This line is indicated in Fig. 72. 

The length of the card between the lines a and b is 
proportional to the length of the engine stroke and there- 




Atmospheric Line ■ 



Fig. 71. 





i 

N 

J 


a 




V\ 




V 


^ 7 Atmospheric Line 














Line of Zero Pressure ^, 






' ' 


i i i i i i ., . i i 





Fig. 72. 



fore to the piston displacement, that is, to the volume 
swept through by the piston. Knowing the clearance 
volume of the engine as a percentage or fraction of the 
piston displacement, this fraction of the length of the 
diagram can be laid off from the end of the diagram to 
give a line of zero volume. This line is also indicated in 
Fig. 72. 

With the line of zero pressure and the line of zero 
volume drawn in, all values of steam pressure and volume 
occupied by steam can be read directly from the diagram, 
and it thus forms a picture of what occurs within the real 
engine cylinder. 

The indicator diagram is used for a number of purposes, 
the more important being: 



120 



STEAM POWEE 



(1) The determination of the energy made available 
within the cylinder, that is, the indicated horse-power, I.h.p. 

(2) The determination of the amount of initial conden- 
sation and of heat interchanges between walls and cylinder. 

(3) The determination of what is known as the diagram 
water rate. 

(4) The study of the operation and timing of valves. 
The second one of these uses has already been considered 
in Chapter VII, the others are treated in succeeding sections. 

64. Determination of I.h.p. The lines of the indicator 
diagram show by their height the pressures or forces acting 
on the engine piston as it moves. But the product of force 
by distance is equal to work and these lines can be used there- 
fore for determining the net 
work done by the steam 
upon the piston. 

In Fig. 73 is shown the 
upper part of the diagram, 
the curved lines represent- 
ing the successive pressures 
in pounds per square inch 
which acted on the left face 
of the piston while it moved 
outward. If the average 
pressure could be deter- 
mined and multiplied by 
the area of the piston face, this product would be the 
average total force acting on the piston. Multiplying this 
by the distance traveled would give the work done by 
the steam upon the piston. Expressed in the form of an 
equation, 

E = poXaXLftAbs., (33) 

in which 

Eo = work done upon piston by steam during outstroke ; 
po = mean pressure (in pounds square inch) acting on 
piston during outstroke ; 





1 


-Length of Diagram s»-» 

>^ in Inchas 1 

illlilv 


1 


a 




1 1 




1} 1 1 


5 






1 ! 
1 1 
1 J 



■:-f^ 



Fig. 73. — Positive Work Area. 



INDICATOR DIAGRAM AND DERIVED VALUES 121 

a = area of piston face in square inches; and 
L = stroke of piston in feet. 

For the instroke shown in Fig. 74, the work done by 
the piston on the steam is given by the similar expression m 



Ei=piXaXL ft.-lbs., 



(34) 




Fig. 74. — Negative Work Area. 



in which Ei and p% represent work done and mean pressure 
respectively. 

The net work done by the steam upon the piston per 
cycle is then, 

# C ycie =E -Ei= (p - Pi) oL f t.-lbs. . . . (35) 

The values of po and pi can be found directly from the 
diagram by dividing the areas A and A t respectively by the 
length I and then multiplying by the scale of the spring, 
giving 

A 



Po 



and 



I 
A* 



X scale of spring, 



p* = -7-Xscale of spring, 



122 STEAM POWER 

so that, 

Po-Vi = V = — ] — *X scale of spring . ... (36) 

= areaof f agram X S cale of spring. . (37) 

The value of p evidently can be determined very simply 
from the indicator diagram, and the work per cycle can be 
found when p is known by substituting in the following 
equation, obtained by putting p for po — pi in Eq. (35), 

# C ycie = ;pXaXLft.-lbs (38) 

The pressure p is known as the mean effective pressure and 
is often represented by M.E.P. 

If n cycles are produced per minute, the net work done 
by the steam upon the piston per minute will be 

E m m = pXaXLXn, (39) 

which is generally rearranged to read, 

E m [n=pLan, (40) 

in which form the group of letters forming the right-hand 
member is easily remembered. 

Since 33,000 foot-pounds per minute are equivalent 
to one horse-power, it follows that the power made avail- 
able as shown by the indicator diagram, that is, the indicated 
horse-power, must be, 

T , pLan N 

Lh - p - = 337W (41) 

in which 

p = mean effective pressure in pounds per square inch ; 

L = stroke of piston in feet ; 

a = area of piston in square inches; 

= (diam. cyl. in inches) 2 X7r/4 = .7854d 2 ; and 
n = number of cycles per minute. 



INDICATOR DIAGRAM AND DERIVED VALUES 123 

If an engine cylinder takes steam on one side of the 
piston only, that is, if the cylinder is single acting, the num- 
ber of cycles produced per minute is equal to the number 
of revolutions per minute, but it should be noted that for 
other arrangements this is not necessarily true. In the 
case of double-acting engines which receive steam at both 
ends of the cylinder, the number of cycles produced is equal 
to twice the number of revolutions. 

It should also be noted that the symbol a represents 
the area of the piston face upon which the steam acts. 
If a piston rod extend from the face of the piston to and 
through the cylinder head (as is always the case at the 
crank end of double-acting cylinders), the area of the 
piston rod must be subtracted from that of the piston to 
obtain the area on which the steam really acts. When a 
tail rod is used, a correction must be made for each side 
of the piston. 

In the case of double-acting engines the indicated 
horse-power may be determined in two ways: It may be 
figured separately for the two ends of the cylinder, or 
the values for the area and pressure may be averaged for 
the two ends and the value of n chosen equal to twice 
the revolutions per minute. The former is generally the 
more accurate method. 

It will have been observed that the area of the indi- 
cator diagram must be determined before the mean 
effective pressure can be found. This area is generally 
measured by means of an instrument known as a planimeter, 
and this is the most accurate method. It occasionally 
happens, however, that a planimeter is not available when 
the value of the indicated horse-power is desired. Under 
such circumstances an approximate determination of the 
area of the indicator diagram can be made by the method 
of ordinates. 

For this purpose the length of the diagram is divided 
into an equal number of parts, usually ten, as shown in 



124 



STEA.M POWER 



Fig. 75 and vertical lines y\, 2/2, 1/3, etc., are drawn at the center 
of each of the parts into which the diagram has been divided. 
The mean ordinate or height is then found from the equa- 
tion, 

= 2/i+?/2+?/3+?/4+etc. 
number of vertical lines' 



(42) 



and the mean effective pressure is then determined by 
multiplying y m by the scale of the spring. 

An indicator diagram similar to that shown in Fig. 76 
is occasionally obtained. The small loop on the end repre- 
sents negative work, since the pressure of the steam which 




w m m 




Fig. 75. 



Fig. 76. 



does work upon the piston is lower than that which resists 
the return of the piston. When using a planimeter, this area 
is automatically subtracted from that of the rest of the 
diagram, but care should be taken to see that this is also 
done when the method of ordinates is used. 



ILLUSTRATIVE PROBLEM 

1. Determine the I.h.p. of a double-acting steam engine, having 
a cylinder 8 ins. diameter, length of stroke, 12 ins., running at 
100 R.P.M., the mean effective pressure (M.E.P.) on the piston 
being 45 lbs. Neglect the area of the piston rod. 



I.h.p. 



pLan _(pXa) lbs.X(Ln) ft. per min. 
33,000 = 33,000 ft.-lbs. per min. 
(45X8X8X.7854) lbs. XffX 100X2) ft. per min. 
33,000 



INDICATOR DIAGRAM AND DERIVED VALUES 125 

2260 lbs. X 200 ft. per min. 
33,000 

= 14 nearly. 

2. The I.h.p. of a double-acting engine is 14, the R.P.M. =100; 
M.E.P. =45 lbs.; length of stroke = 12 ins. Find the diameter 
of the cylinder, neglecting area of piston rod. 

First determine the area of the piston from the formula 

vLan 33,000 I.h.p. 

Lh - P '=33^0 ° r "' pXLXn ' 

33,000X14 „ A . rf 
° = 45XlX100X2 =51 - 4sq - m - = T ; 



,7 I 51 - 4 



— - '—- = Vq5A=S ins. (approx.), 

7854 



65. Conventional Diagram and Card Factors. It is 

often necessary to approximate the mean effective pressure 
obtained in the cylinder of an engine when no indicator 
diagrams are available. The most common case is when 
an engine is being designed to carry a certain load and it 
is desired to determine the necessary cylinder dimensions 
and speed. If the probable mean effective pressure can be 
determined, the dimensions and speed can be found from the 
equation, 

pLan 



I.h.p. per cylinder end = 
I.p.h. 



33,000' 
by rewriting it 

pLnX0.7854:d 2 
33,000 
from which 



d? 33 000I.hp. 

0.7854 pLn v y 

Since n is equal to revolutions per minute for one cylinder 
end, the product of L by n must be equal to half the piston 



126 STEAM POWER 

speed of the engine, and a proper value of this product can 
be chosen for substitution in the equation. If a proper 
value for p can then be predicted the only unknown remain- 
ing will be the diameter d, and this can be found by solving 
the equation. 

The prediction of the mean effective pressure is made 
either by drawing upon recorded experience in the form 
of values obtained in similar engines previously constructed 
or by means of what is known as a conventional indicator 
diagram. 

The conventional diagram is drawn with upper and 
lower pressures equal to those expected in the case of the 
real engine, and all expansions and compressions are drawn 
as rectangular hyperbolas. The equation of the rectangu- 
lar hyperbola is 

PiV^PnVn, (44) 

in which subscript 1 indicates initial conditions and sub- 
script n represents any later conditions with the same mate- 
rial in the cylinder. This law is assumed because it is the 
simplest and, as a rough average, gives values as close to 
those actually attained as do any of the more complicated 
laws. 

The diagram may be drawn as nearly as possible like 
the one which the engine may be expected to give or it 
may be drawn with various simplifications which remove 
it more and more from the approximation to an actual 
indicator diagram. In any case, the mean effective pres- 
sure is determined from this diagram and this value is then 
multiplied by a corrective factor, the value of which has 
been determined by experience. This corrective factor is 
called the diagram factor or card factor and it is realty 
the ratio of the area of the diagram the engine would 
really give to the area of the conventional diagram 
used. 

The simplest form of conventional diagram is drawn 



INDICATOE DIAGRAM AND DERIVED VALUES 127 



by neglecting the clearance volume and has the shape shown 

in Fig. 77. The upper line is drawn horizontal at a height 

representing the highest pressure expected and of such a 

length (compared with the length of the diagram) as will 

approximately represent the fraction of the stroke at which 

cut-off is to occur in the real 

engine. The expansion curve is 

then drawn in as a rectangular 

hyperbola and extended until the 

end of the diagram is reached. 

The next line is drawn vertical 

and the lower line of the diagram 

is drawn horizontal at a height 

representing the pressure expected 

in the space into which the engine 

is to exhaust. 

This simple diagram can be 
divided into the three areas shown 

and the value of the work represented by these areas can be 
determined from the equations given below, the first and 
last of which should be self-evident from what has preceded. 
The equation for the work represented by area A 2 can be 
determined very easily by means of integral calculus. The 
equations are, 




Fig. 77.— Conventional In- 
dicator Diagram. 



and 



Ai represents P\V\ ft. -lbs.; 

A2 represents P\V\ log e ^- = Pi~Fi log e r ft.-lbs., 



As represents P2V2 ft.-lbs., 



in which P represents pressure in pounds per square foot and 
V represents volume in cubic feet. 

The total area is then equal to the sum A1+A2 — As 
and the net work is equal to a similar sum of the right- 



128 STEAM POWER 

hand members given above. The net work must also 
equal the mean-effective pressure P m multiplied by the 
total volume change, so that 



and 



PmV 2 = PiVi+PiV 1 \oger-P 2 V2, . . (45) 
P m =Pi^+Piploger-P 2 . . . (46) 

^(ft+ft 10 ^) - ^' • • • (47) 

1 Vi . . 
and substituting - for ■=- this gives 
r v 2 

l±^_p 2 (48) 



Pm = P 



v 2 

The ratio T^=r is called the ratio of expansion and its 

Vi 1 
reciprocal, ■==- = — is known as the cut-off ratio. By the use 

of this ratio the volume terms can be disposed of and the 
equation above is obtained. This equation then gives the 
mean effective pressure in terms of upper and lower pres- 
sures and the fraction of the stroke at which cut-off is 
desired in the real engine and no cylinder dimensions need 
be known. 

Since pressures in steam-engine practice are usually 
given in pounds per square inch, the equation for mean 
effective pressure is more useful in the form 

Pm = J l +^l)- P2 , .... (49) 



in which pi and p 2 and p m are expressed in pounds per 
square inch absolute. For convenience in the use of this 



INDICATOR DIAGRAM AND DERIVED VALUES 129 

equation the values assumed by the bracketed quantity 
are given for various conditions in Table III. 



TABLE III 



r 


1 +log e r 
r 


r 


1 +log e r 
r 


r 


1 +log e r 
r 


1.0 


1.00 


6.0 


0.465 


16.0 


0.236 


1.5 


0.937 


7.0 


0.421 


17.0 


0.226 


2.0 


0.847 


8.0 


0.385 


18.0 


0.216 


2.5 


0.766 


9.0 


0.355 


19.0 


0.208 


3.0 


0.700 


10.0 


0.330 


20.0 


0.200 


3.5 


0.644 


11.0 


0.309 


21.0 


0.192 


4.0 


0.597 


12.0 


0.290 


22.0 


0.186 


4.5 


0.556 


13.0 


0.274 


23.0 


0.180 


5.0 


0.522 


14.0 


0.260 


24.0 


0.174 


5.5 


0.492 


15.0 


0.247 


25.0 


0.169 



The values of the mean effective pressures obtained 
from this form of diagram are very much higher than are 
to be expected from real engines with the same initial and 
terminal pressures and the same nominal ratio of expan- 
sion. They are therefore corrected by multiplying by the 
proper diagram factor as selected from Table IV. It is 
obvious from the range of values given that the selection 
of a proper value for the factor depends largely on expe- 
rience, but such experience is quickly gained by contact 
with real engines and a study of the practical diagrams. 

TABLE IV 
Diagram Factors 

Simple slide-valve engine 55 to 90% 

Simple Corliss engine 85 to 90 

Compound slide-valve engine 55 to 80 

Compound Corliss engine 75 to 85 

Triple-expansion engine 55 to 70 

66. Ratio of Expansion. — The ratio of expansion used 
above is sometimes called the apparent ratio. It is not the 



130 



STEAM POWER 



real ratio of expansion for an engine with clearance, 
such an engine the real ratio of expansion is 



For 



W + Va 



(50) 




Fig. 78. 



in which the symbols represent 
the volumes indicated in Fig. 
78. 

The numerical values of r 
and r' are often very different 
and care should be used in dis- 
tinguishing between them. The 
diagram factors referred to in 
Table IV are for idealized con- 



ventional cards without clearance as shown in Fig. 77. 



ILLUSTRATIVE PROBLEMS 

1. Given an engine with a stroke of 24 ins. and cut-off occurring 
at \ stroke. Steam pressure of 160 lbs. per square inch and 
back pressure of 16 lbs. Assume diagram factor = 80%. Neglect- 
ing clearance, find the probable M.E.P. 



M.E.P. =p 



l+log e r 



p 2 = 160 



l+log e 3 



r / \ 3 

<.7-16=112.0-16=96 1bs. 



16 



Hence probable M.E.P. =.80X96 =76.8 lbs. 

2. A given double-acting engine indicates 75 I.h.p. under the 
following conditions: 

Cut-off at 20%; steam pressure, 140 lbs. per square inch 
absolute; piston speed, 600 ft. per minute; back pressure, 2 lbs. 
per square inch absolute. 

Assume a diagram factor for this type of engine equal to 85% ; 
and neglecting clearance, find a convenient size of the cylinder 
(diameter and stroke). 



INDICATOR DIAGRAM AND DERIVED VALUES 131 

Solution. 

r= Jo =5; 



^ 1+bger j 



/ l+Iog e 5 \ 
73.1 -2 =71.1 lbs. per sq.in. 



Diagram factor =85%. Hence probable 
M.E.P.=71.1X.85 =60.4 lbs. 

Therefore, since 

T , pLan 75X33,000 co n . 

Lh - P - = 3P00 ft= 6O«600- = 68 - 3Sq - m - 

d=9% ins. (approx.); 
and since 2Ln = 600, assume L = 1 ft. 

hence n=300R.P.M. 

The engine is rated 9.5X12 ins., running at 300 R.P.M. 

67. Determination of Clearance Volume from Diagram. 

It was shown in a preceding paragraph that the clearance 
volume of a cylinder must be known in order to draw the 
line of zero volumes on the indicator diagram. This 
volume can be determined accurately for any real engine 
by weighing the quantity of water required to fill the clear- 
ance space, but this procedure is often impossible and 
an alternative, though approximate, method is often 
resorted to. 

This method is graphical and depends upon the assump- 
tion of the law of expansion and compression. As in the 
case of the conventional diagrams, expansion and compres- 
sion are assumed to follow rectangular hyperbolas. 

It is a property of this curve that diagonals such as 
aa and bb drawn for rectangles with their corners on the 



132 



STEAM POWER 



curve all pass through the origin of coordinates as shown 
in Fig. 79. 

If two points a and c are selected on the expansion 
curve of a real diagram and a rectangle is drawn upon 
them as shown in Fig. 80, the diagonal bd extended will 
pass through the origin of coordinates, if the expansion 
follows the assumed law. The point at which this diagonal 
cuts the zero pressure line must therefore be the point 
through which the vertical line of zero volume is to be drawn. 

If the original assumption were correct, this construc- 
tion would give the same point when different locations 
of the points a and c were chosen and when used on the 





Fig. 79.— Rectangular Hyperbola. 



Fig. 80. 



compression as well as on the expansion line. In reality- 
it will generally give as many different locations for the 
origin as are chosen for the rectangle abed. It is customary 
to construct this rectangle of fair size and to locate it near 
the center of the expansion curve. 

68. Diagram Water Rate. When an engine is run on 
saturated steam, part of the steam supplied is generally 
condensed upon the cold metal walls surrounding it. The 
indicator diagram therefore shows the volumes assumed by 
the mixture of steam and liquid water in the cylinder, but, 
since the volume occupied by the liquid is negligible, it 
may be assumed to show the volumes occupied by the part 
of the mixture which exists in vaporous form. 



INDICATOR DIAGRAM AND DERIVED VALUES 133 

Assuming that the vapor is saturated, the volume 
occupied by one pound at various pressures can be found 
from the steam tables and, therefore, the weight existing 
in the cylinder can be calculated. The weight of steam 
determined in this way is known as the indicated steam, 
the diagram steam or the diagram water rate. 

The diagram water rate is generally determined for a 
point such as z in Fig. 81 just 
after cut-off, though some 
engineers prefer to use a point 
nearer the lower end of the 
expansion curve. The volume 
occupied by the steam con- 
tained in the cylinder at point 
z is equal to V z and its weight 
can be determined by dividing 
this volume by the specific volume V 2 for the existing 
pressure P z . Thus, calling the weight of steam in the cyl- 
inder w z , 

V z 




Fig. 81. 



w >=vl- 



(51) 



This quantity of steam is a mixture of cylinder feed 
and clearance or cushion steam and the weight of the latter 
must therefore be subtracted from w z to obtain the weight 
of cylinder feed w f . Assuming the cushion steam dry and 
saturated at the point k, the weight of cushion steam is 



Wt 



V*' 



(52) 



so that the weight of cylinder feed per cycle as shown by 
the diagram at the point z is 

Wf=W z -W t = y- y~. ...... (53) 



The formula is generally modified to give the steam 
consumption per indicated horse-power hour, instead of 



134 STEAM POWER 

per cycle, and it is also expressed in different terms as a 
matter of convenience. 
For this purpose let 

ya = clearance volume divided by piston displacement per 
stroke 

lei 

~v 

2/ z = piston displacement to point z divided by piston dis- 
placement per stroke 

Jl 

V 

yt = piston displacement to point k divided by piston dis- 
placement per stroke 

J± 
V 
a = area of piston in square inches; 
p=mean effective pressure in pounds per square inch; 
L = stroke in feet ; and 
n = number of cycles per minute. 

The piston displacement is then tj^L cubic feet and 

the volumes at z and k are given by 



and 



'*Xffi) + (v«xj§ 



v>xm)+(y*xm): 



Substituting these values in Eq. (53) gives the indicated 
cylinder feed per cycle as 



w f 



aL / yz+yci y k -hyci \ , . 

= m\~T 2 vr) (54) 



Multiplying by the number of cycles per hour (60 Xn) 

and dividing by the indicated horse-power, }* gives 

00, uuu 



INDICATOR DIAGRAM AND DERIVED VALUES 135 



the diagram water rate, or steam shown by the diagram per 
I.h.p. hour as 



Wa 



13,750 

V 



•■+ya yt+y a 
v 2 v t 



). 



(55) 



in which form the equation involves only values which can 
be determined directly from the diagram without any 
knowledge of the engine dimensions. 

The value obtained for w d will vary as the location of 
points z and k are varied because of the quality changes 
occurring during expansion and compression, and it is 
obvious that the diagram water rate is in no sense an accu- 
rate measure of the real water rate of the engine. It is, 
however, often useful for comparison with the real water 
rate, the ratio giving an indication of the loss by conden- 
sation and leakage. 

Average values for real water rates are given in Chapter 
XI. 



ILLUSTRATIVE PROBLEM 

Given the diagram shown in Fig. 82 and the following data 
from an actual test, find the diagram water rate for point c, and 
for point n. Double-acting steam 
engine having: 

Average piston area =28.9 sq.- 
in.; 

Length of stroke = 8 in.; 

R.P.M.=237; I.h.p. =8.75; 
M.E.P. =31.6 lbs.; 

Clearance = 13% ; Beginning of 
compression = 29% ; 

Weight of condensate per hour 
=371 lbs.; 

Quality at throttle =95%; 

Sp. vol. at c=7.8; 

Sp. vol. at n = 12.57; 

Sp. vol. at K =38.4 cu.ft. per lb.; 

Assume xt = 100%. 



& ,_,_ 




-4 — i-V 


i 


—r\-fc--i— 1 


■ Ti 


!.:•»! ;f v 



Fig. 82. 



ia6 STEAM POWER 

Solution. Substitution in Eq. (55) gives 
13,750 ftjc+yci yt+ya 

m)c= ^~\~v c vT 

13,750/ 0.39+0.13 _0. 29+0.13 \ 
" 31.6 \ 7.8 38.4 / 

= 24.2 lbs. per I.h.p. per hour at point c. 

13,750 fyn+ya yt+ya 

( ^ = ~3ix r^r~~^r 

13,750 /0.638+0.13 _ 0.29+0.13 \ 
" 31.6 \ 12-57 ~ 38.4 / 
= 21.83. lbs. per I.h.p. per hour at point n. 

371 
Real water rate =——X 0.95 =40.2 lbs. 

8.75 

69. T^-diagram for a Real Engine. In Chapter VI 
the T ^-diagrams of the various ideal cycles were given and 
attention was called to the fact that these diagrams were 
particularly useful, because they showed certain things 
which were not apparent from the more common PV- 
diagrams. 

It has been customary for many years to draw T<p- 
diagrams for real engines by " transferring " the PV- 
diagram to T^-coordinates, and various analytical and 
graphical methods have been developed for this purpose. 
There are certain unavoidable errors in all the methods 
used for drawing these diagrams, and the expansion curve 
is the only one of all the lines finally obtained which has 
any claim to accuracy. Even this curve is generally incor- 
rectly interpreted, because a knowledge of the exact weight 
of clearance steam is necessary for an accurate interpreta- 
tion and such knowledge is never available. 

Under the circumstances it seems unnecessary to con- 
sider in this book the rather complicated details involved 
in the construction of T^-diagrams purporting to show 
the behavior of steam in real engines. 



INDICATOR DIAGRAM AND DERIVED VALUES 137 

70. Mechanical and Thermal Efficiencies. The method 
of obtaining the indicated horse-power from the indicator 
diagram has been given in preceding paragraphs. In the 
real engine this power is not all made available at the shaft, 
because some of it is used in driving the engine against 
its own frictional resistance. Calling the power lost in 
this way the friction horse-power, it follows that 

I.h.p. =F.h.p.+D.h.p, (56) 

in which 

I.h.p. = indicated horse-power determined from the real 

indicator diagram; 
F.h.p. = friction horse-power, i.e., power required to 

drive engine; and 
D.h.p. = developed horse-power, i.e., power made avail- 
able at shaft. This is sometimes called brake 
horse-power and abbreviated B.h.p. 

The developed horse-power is therefore always less 
than the indicated horse-power. The better the construc- 
tion of the engine the smaller the friction loss, and the 
measure of this loss is usually given in the form of an effi- 
ciency. It is called the mechanical efficiency, and is defined 
by the equation 

Mech. erT.^^^ (57) 

I.h.p. 

The developed or brake horse-power can generally be deter- 
mined experimentally in one way or another so that it 
becomes possible to evaluate numerator and denominator 
of Equation (57) and to determine the mechanical efficiency 
of the engine. Methods of determining the developed 
horse-power are considered in Chapter XX. Values of 
mechanical efficiency range from about 80 per cent in the 
case of poorly designed and poorly adjusted horizontal 
engines to about 95 per cent in the case of the best vertical 
designs. 



138 STEAM POWER 

The efficiency determined by dividing energy made avail- 
able by heat supplied is known as the thermal efficiency. 
There are two possible thermal efficiencies, one based on 
the indicated power and the other on the developed power. 
The former is called the thermal efficiency on the indicated 
horse-power or the indicated thermal efficiency; the other 
is known as the thermal efficiency on the developed horse- 
power or the developed thermal efficiency. Obviously 

Dev. ther. eff . = Mech. eff X Indie, ther. eff. . (58) 

The heat supplied may be assumed in two different 
ways; it may be taken as the total heat above 32° F. in the 
steam supplied the engine, or it may be taken as this value 
less the heat of the liquid corresponding to exhaust tem- 
perature. The second method is preferable, since it is 
reasonable to assume that the exhaust steam can be con- 
densed to water at the same temperature or raise an equal 
weight of water to that temperature and that this water can 
be pumped to the boiler with the heat of the liquid corre- 
sponding to this temperature. This is practically parallel 
to the assumption made in treating the theoretical cycles. 

The thermal efficiencies are then 



Indie, ther. eff. 

I.h.p.X2545 



Heat above liquid at exhaust temp, per hour 7 
and 

Dev. ther. eff. 

D.h.p.X2545 

Heat above liquid at exhaust temp, per hour 



(59) 



(60) 



Values of the indicated thermal efficiency range from 
about 5 per cent in ordinary practice with small engines to 
about 25 per cent in the best large engines. Values as low 
as 1 per cent are not uncommon with small, poorly designed 
and poorly operated engines. 



INDICATOR DIAGRAM AND DERIVED VALUES 139 

The actual performance of the cylinder of an engine 
is sometimes compared with the ideal possibilities as indi- 
cated by the Clausius and the Rankine cycles. The ratio 
of the work obtained in the real engine to that which could 
be obtained from the same quantity of heat with a Rankine 
or Clausius cycle is a measure of the performance of the 
real cylinder. This ratio is variously designated as cylinder 
efficiency, indicated efficiency, relative efficiency, etc. Its 
values range from less than 40 per cent to over 80 per cent, 
the highest recorded value being just over 88 per cent. 

PROBLEMS 

1. Using Table I, Chapter I, plot the specific heat of water 
between the range of temperatures of 20° F. and 300° F. for the 
intermediate values given. By the ordinate method for finding 
the mean height of an indicator diagram, determine the mean 
or average specific heat over this range. i 

2. A double-acting engine Is required to give 50 I.h.p. under 
the following conditions : 

Cut-off =25%; 

Steam pressure = 150 lbs. per square inch absolute; 
Back pressure = 16 lbs. per square inch absolute; 
Piston speed =540 ft. per minute. 

If the diagram factor for this type of engine is 75%, find the 
diameter of the cylinder and select the stroke and R.P.M. 

3. Assume a single-acting engine with 10-in. diameter and 12-in. 
stroke, 10X12 ins., to have cut-off occur at various points between 
10% and 50% of stroke. Assume also the pressures, speed, and 
card factor as given in Prob. 2. Find the probable I.h.p. at 
different cut-offs. 

4. Given an 18X24-in. engine running at 120 R.P.M. 
Back pressure =2 lbs. per square inch absolute; 
Clearance = 10%; 

Cut-off =40%; 

Diagram factor =85%. 

Supposing cut-off to remain constant, find the I.h.p. 's cor- 
responding to steam pressure of 50, 90, and 130 lbs. per square 
inch absolute. 

5. Find the weight of dry steam which must be supplied per 
I.h.p. hour for each case of the previous problem, assuming the 



140 STEAM POWER 

quality at cut-off to be 80%. Assume compression pressure to 
be 30 lbs. absolute and that steam is dry and saturated at end of 
compression. 

6. Find the quality of steam at cut-off in a cylinder, in which 
the piston displacement is 0.1278 cu.ft.; clearance = 10% ; cut-off 
at 25% stroke; steam pressure at cut-off, 115 lbs. per square 
inch absolute, and weight of steam in the cylinder at cut-off = 
0.012 lb. 

Note. Quality == T . — '—. r for the given pressure. 

^ J Weight XSp. vol. & F 

7. The piston displacement of a certain engine is 0.2 cu.ft. 
What weight of steam is in the cylinder at release where quality 
is 90%, and pressure is 25 lbs. per square inch absolute, if the 
clearance is 10%, and release occurs at 95% of the stroke? 

8. Find the weight of cushion steam in a 6X6 in. engine in 
which clearance = 15%; compression begins at 85% of the return 
stroke; back pressure is 14.7 lbs. per square inch absolute, and 
the quality of the cushion steam at the beginning of compression 
is 95%. 

9. Find the pressure and quality at the end of the compression 
line of the previous problem, assuming it to be adiabatic. 

10. An 8X10 in. engine running at 300 R.P.M. is double-acting, 
and cuts off at 15% of the stroke at a pressure of 120 lbs. per 
square inch absolute. It has a steam consumption of 35 lbs. per 
I.h.p.-hour. The compression begins at 60% of the return stroke 
with a quality of unity and a back pressure of 5 lbs. per square 
inch absolute. Clearance = 10%. 

If this engine delivers 27 H.P., and has a mechanical efficiency 
of 90%, what is the quality at the point of cut-off? 

11. In the previous problem, assume release to occur at 90% 
of the stroke with an absolute pressure of 30 lbs. per square inch. 
What is the quality at this point? 

12. A certain engine gives one horse-power hour at the shaft 
for every 20 lbs. of steam supplied. The steam has an initial 
pressure of 150 lbs. absolute and is dry and saturated when it 
arrives at the engine. The back pressure against which steam is 
exhausted is 4 lbs. absolute. 

(a) Find the thermal efficiency of this engine on the developed 
or shaft horse-power. 

(6) If the mechanical efficiency of the engine is 90%, what 
is the value of the thermal efficiency on the indicated horsepower? 



CHAPTER IX 



COMPOUNDING 




Voliame 



71. Gain by Expansion. The cycle which gives a 
rectangular PF-diagram is the least economical of all the 
ideal cycles described in Chapter IV. This comes from 
the fact that none of the heat stored in the steam is con- 
verted into work when this cycle is used. Thus, if the 
cylinder shown by full lines in 
Fig. 83 operate on this cycle 
and be of such size that it will 
receive just one pound of steam 
per cycle, it makes available an 
amount of work represented by 
the area abed. The positive 
work done by the steam upon 
the piston is the equivalent of 
the external latent heat of 
vaporization while no use is 
made of the heat stored in the 
steam. This stored heat is re- 
moved as heat during the con- 
densation and exhaust, which give the lines be and cd. 

If a piece be added to the cylinder as indicated by 
the dotted lines, the same quantity of steam will make more 
heat available by expanding after cut-off, as shown by the 
curve be, the net work in this case being represented by 
the area abefd instead of by the smaller area abed. But 
the heat supplied is the same in both cases, namely that 
required to form one pound of steam at the pressure Pi, 
so that the use of a large cylinder and the incomplete ex- 

141 



i n 

i 
i 

i i 
i i 

I ! 
I 



Fig. 83. 



142 



STEAM POWER 



pansion cycle results in the development of more work 
than can be obtained with the rectangular cycle from the 
same amount of heat. 

Obviously it would be theoretically advantageous to 
add still more to the length of the cylinder and allow the 
expansion to continue to back pressure, giving the com- 
plete expansion cycle as shown in Fig. 84, thus obtaining 
the maximum quantity of work at the expense of the heat 
stored in the steam supplied the cylinder. Practically, 
it is found inadvisable to continue the expansion to such a de- 




Fig. 84. 



gree in reciprocating steam engines, because at low pressures 
the volume increases very rapidly for small pressure drops. 
Thus a great increase is necessary in the size of the cylinder 
if the last part of the expansion is to be completed, but 
the amount of work obtained is comparatively small, as 
shown by the small height of the long toe thus added to 
the diagram. This may result in an actual loss, because the 
increased friction losses of the very large cylinder may 
more than balance the small increase of net work gained 
by its use. It thus results that, in every real reciprocating 
engine, there is some point beyond which it is not economi- 
cal to carry the expansion, and the incomplete expansion 



COMPOUNDING 143 

cycle is therefore approximated in such engines rather 
than the cycle with complete expansion. 

Viewing the matter from another angle, a cylinder of 
a certain size may be assumed as shown in Fig. 85. The 
use of the rectangular cycle 
abed in this cylinder will make 
available the maximum quan- 
tity of work possible with the 
upper and lower pressures 
chosen. If cut-off be made 
to occur earlier as at b' ', the 
expansion b'c' will result in a 
loss of the quantity of work 
obtained, as shown by the 
area b'bc' ', but the steam used yig. 85. 

per horse-power will be less, 

so that there will be a gain in steam economy. Putting the 
cut-off still earlier will cause a still greater loss of work 
obtained from a cylinder of the chosen size, but theoreti- 
cally will result in greater economy of steam. 

Summing up, it may be said that the greater the ratio 
of expansion the greater should be the economy in the use 
of steam on a theoretical basis. 

The lower pressure is set in real engines by the pressure 
in the space into which the engine is to exhaust. If the 
engine is to be operated non-condensing, the atmospheric 
pressure determines the lowest possible exhaust pressure; 
if the engine is to be operated condensing, the exhaust 
pressure is set by the lowest pressure which can be eco- 
nomically maintained in the condenser. 

There is thus a real limit to the extent to which expan- 
sion can be carried in any real engine with a given initial 
pressure. A certain drop must exist at the end of the 
diagram, for reasons already explained, and an expan- 
sion line drawn backward from the top of the line repre- 
senting this drop will give the earliest possible cut-off 



144 STEAM POWER 

which can be used in the engine with a given initial 
pressure. 

The ratio of expansion can be further increased, how- 
ever, by raising the initial pressure as indicated by the 

dotted lines in Fig. 86, 
P 



Fig. 86. 



and the limit in this direc- 
tion would come with the 
inability of materials of 
construction to withstand 
the resulting strains. 

These conclusions 
drawn from the facts 
V -J t developed above must all 
be modified in the case of 
real engines, because of 
the effect of cylinder condensation. This has been shown 
to increase as the cut-off is made earlier and as the 
pressure (and therefore the temperature) range in a cyl- 
inder is increased. There is, therefore, a limit beyond 
which it is not advisable to carry the ratio of expansion in 
a single cylinder. 

Experience has shown that in the case of engines expanding 
steam entirely in one cylinder (simple engines) the best com- 
mercial results are obtained when (a) they are operated non- 
condensing, (b) the initial pressure is between 80 and 100 lbs. 
per square inch for the simpler forms of valves and below 125 
lbs. for the better forms, and (c) the point of cut-off is at about 
i stroke with the simpler valves and at % to \ stroke with the 
better forms of valves. Strictly, these statements apply only 
to engines of the types so far treated. The Una-flow engine, 
described later, has entirely different characteristics. 

72. Compounding. If the ratio of expansion of ordinary 
types is to be increased above the values just given, some 
means must be used for the reduction of loss by con- 
densation. This loss can be reduced by decreasing the 
surface exposed to high-temperature steam and by de- 



COMPOUNDING 



145 



creasing the temperature range in a cylinder. Both of 
these results can be achieved by what is known as 
compounding. 

Assume that it is deemed advisable to produce a cycle 
similar to that shown in Fig. 87 (clearance neglected) and 
that in order to obtain 
high steam economy (low 
water rate) the ratio of 
expansion chosen is very 
much greater than four. 
No gain in economy 
would result from such 
excessive expansion in a 
single cylinder, in fact 
there would be a well- 
defined, unavoidable loss. 
But suppose that the high- 
pressure steam is admitted 
to a small cylinder such 
as that shown and is ex- 
panded to the point /, is 
then exhausted as shown 
by fg into the larger cyl- 
inder along gf and then expanded to the point c in the 
larger cylinder. The cycle produced is the same as that 
which would have been obtained by expanding entirely in 
one cylinder, but the surface of the clearance space of 
the high-pressure (H.P.) cylinder, which is exposed to high- 
pressure steam is smaller than it would be in a cylinder of 
the size required to hold the steam when fully expanded and, 
moreover, the lowest temperature to which it is subjected is 
that corresponding to the pressure at / instead of the 
much lower temperature corresponding to the pressure 
at d. 

The condensation which would occur in the H.P. cylinder 
would obviously be less than that which would result from 




Low Pressure Cylinder 
L.P. Cylinder 



Fig. 87. 



146 STEAM POWER 

the use of one large cylinder and, remembering that the 
greater part of the heat given up during condensation is 
received again by the steam during exhaust, it is obvious 
that approximately this same quantity of heat can again 
be given to the low-pressure cylinder walls. Thus, although 
there are two cylinders in which condensation and re- 
evaporation occur, and although the sum of the heat given 
to the walls of the high-pressure cylinder and that given 
to the walls of the low-pressure cylinder might be greater 
than that given to the walls of a single cylinder under similar 
conditions, the use of two cylinders results in a consider- 
able saving because loss in the high-pressure cylinder is 
practically wiped out by the exhaust of the heat concerned 
into the low-pressure cylinder. 

If the loss by radiation and conduction from the high- 
pressure cylinder be neglected, the result of the use of two 
cylinders is practically to limit the loss by condensation 
and re-evaporation to that occurring in the low-pressure 
cylinder. As the ratio of expansion in this cylinder is in 
the neighborhood of that common in simple engines, or 

even less, and as the tem- 
perature range is small, 
the net loss is also small. 
It is obvious that the 
smaller the surface of the 
high-pressure cylinder can 
be made, and the smaller 
the temperature range in 
a single cylinder, the 
smaller will be the net loss 
Fig. 88. by cylinder condensation 

and re-evaporation. A 
saving should therefore be effected by using more than two 
cylinders, and it is not inconceivable that five or more might 
\>e used. The result of using five cylinders is shown in Fig. 
88, and it is evident that the clearance surfaces exposed to 




COMPOUNDING 147 

high temperatures, the temperature ranges per cylinder 
and the ratios of expansion per cylinder are all small. 
The gain in econ omy should therefore be correspondingly 
great. 

There are two limits to the possible multiplication of cylin- 
ders in this way. 

(1) As the number increases the radiating surface and 
therefore the heat lost by radiation increases. The extent 
of this effect can be appreciated by noting that every 
cylinder with the exception of the low-pressure cylinder is 
really an unnecessary addition, because the cycle could be 
produced entirely in the low-pressure cylinder. On the 
other hand, the surfaces of cylinders which operate at high 
temperature are small as compared with that which would 
be exposed to this temperature if the entire cycle were pro- 
duced in the low-pressure cylinder. 

(2) As the number of cylinders is increased, the first 
cost, the complexity and the cost of lubrication and attend- 
ance are all increased so that, for each installation, some 
number will be found beyond which the interest on the 
investment and the added cost of operation and mainte- 
nance would more than balance the saving of fuel. 

The second limit mentioned is the more important 
commercially, as it is the first one reached. For ordinary 
operating conditions in stationary power plants expansion 
in two cylinders generally gives the most economical results. 
The total ratio of expansion is generally between 7 and 16, 
that is, the volume of steam at release in the L.P. cylinder 
is from 7 to 16 times the volume at cut-off in the H.P. 
cylinder. For large pumping stations and large marine 
installations, expansion in three cylinders is generally 
considered the most economical, and total ratios of expansion 
of 20 or more are used. Four and five cylinders have 
been used, but the resultant gains do not seem to warrant 
any extensive installation of such units. 

Engines using more than one cylinder for the expansion 



148 STEAM POWER 

of steam in the way just described are called multi-expan- 
sion engines, or compound engines, and the use of multi- 
expansion- is spoken of as compounding. Custom has 
almost confined the use of the term compound engine 
to those in which only two cylinders are used in series as 
indicated in Fig. 89, and such engines are often spoken of 
as 2x engines. 

Engines in which three cylinders are used in series 
are called triple-expansion or 3x engines. With four and 
five cylinders in series the engines are known as quadruple 
or 4x and quintuple or 5x, respectively. 

In the case of triple-expansion engines of large size, 

H.P. Exhaust and L.P. Admission 





To 
Condenser 

Fig. 89. Fig. 90. 

the volume of the low-pressure cylinder required generally 
becomes so great that it is found economical to use two 
low-pressure cylinders instead of one. The flow of steam 
in such an engine is represented diagrammatically in Fig. 
90. This type is known as a four-cylinder, triple-expansion 
engine. 

All multi-expansion engines are generally operated 
condensing, and the choice of type is determined partly by 
the character of work to be done and partly by economical 
considerations. In all cases the boiler pressure must be 
chosen to suit the type of engine used. The pressures 
ordinarily used with the different types are given in 
Table V. 



COMPOUNDING 149, 

TABLE V 
Boiler Pressure Commonly Used 



Type of Engine. 


Boiler Pressure. 
Pounds per Sq.in. Gauge. 


Simple 


80 to 125 


High-speed compound 


100 to 170 


Low-speed compound 


125 to 200 


Triple expansion and higher 


125 to 225 



73. The Compound Engine. The term compound 
engine will be used hereafter in the commercial way as 
reierring to a 2x engine. Such engines may roughly be 
divided roughly into two types, receiver and non-receiver 
engines. The latter are often called Woolf engines, after the 
man who first used this construction. 

A receiver engine has a vessel known as a receiver 
located between the two cylinders and so connected with 
them that the high-pressure cylinder exhausts into the 
receiver and the low-pressure cylinder draws its steam 
from the receiver. By using a receiver the cylinders are 
made independent of each other so far as steam events 
are concerned; the high-pressure cylinder can exhaust at 
any time without reference to the events occurring in the 
low-pressure cylinder. 

A Woolf type has practically no receiver, the high-pres- 
sure cylinder exhausting directly into the low-pressure 
cylinder through the shortest convenient connecting pass- 
age. As the high-pressure cylinder must exhaust directly 
into the low-pressure cylinder it follows that cut-off must 
not occur in the latter until compression starts in the 
former; i.e., very near the end of the stroke. 

An engine with a receiver of infinite size would give 
a horizontal exhaust line for the high-pressure cylinder 
and a horizontal admission line for the low-pressure cylinder, 
since the small amount of steam given to or taken from the 



150 



STEAM POWER 



receiver would have no appreciable effect upon the pressure 
within that vessel. Neglecting throttling losses, the high- 
pressure and low-pressure cards would therefore fit together 
as originally indicated in Fig. 86. 

With receivers of finite size there are pressure changes 
during exhaust by the high- and admission to the low-pres- 
sure cylinders, and real valves and connections also cause 
certain throttling losses, so that the lines representing 
these events are not horizontal nor do they exactly coincide. 

A diagrammatic arrangement of the Woolf engine 
is given in Fig. 91 with idealized diagrams obtained by 




Fig. 91. 



assuming hyperbolic expansions, no clearances, and no 
throttling losses. The pistons must make their strokes 
together in such engines, but they may move in the same 
direction, as shown in the figure, or in opposite directions. 

The ideal diagram would be that shown at (a) by the 
lines AbcdCDA. The idealized high-pressure diagram is 
abcda and the idealized low-pressure diagram is ABCDA. 
The exhaust line da of the high-pressure diagram and the 
admission line BC of the low-pressure diagram are pro- 
duced at the same time. Corresponding points on these 
two lines represent the common pressures assumed by the 
steam not yet exhausted from the high-pressure cylinder, 
the steam in the small connecting passage and the steam 



COMPOUNDING 



151 



already admitted to the low-pressure cylinder. As the 
movement of the low-pressure piston opens up volume 
faster than the high-pressure piston closes up volume, 
the volume occupied by the steam continues to increase as 
the low-pressure piston moves out, and its pressure there- 
fore decreases. 

The two diagrams are shown back to back at (b) in the 
figure and the horizontal line xX connects corresponding 




JTrom 
Boiler 



Fig. 92. 



points on the exhaust of the high pressure and the admission 
of the low pressure cylinders respectively. 

Compound engines are also divided into two types 
on the basis of cylinder arrangement. When the axes of 
both cylinders coincide as shown in Fig. 92 they are called 
tandem compounds. When the axes are parallel as shown 
in Fig. 89, the engines are spoken of as cross-compound 
engines. 

74. Cylinder Ratios. The idealized diagrams of a com 
pound engine with infinite receiver 
volume are shown in Fig. 93 by 
abed and ABODE. The height of 
the high-pressure exhaust line is the 
same as that of the low-pressure 
admission line and represents the 
receiver pressure j)r. The value of 
the receiver pressure is determined 
by the point chosen for cut-off in the low-pressure cylinder. 
Thus if cut-off in the low-pressure cylinder is made to occur 
earlier, as at some point C", the admission line for this 




Fig. 93. 



152 STEAM POWER 

cylinder must move up to B'C and the receiver pressure 
must rise correspondingly. The exhaust pressure in the 
high-pressure cylinder would also rise an equal amount. 

Changing the point of cut-off in the low-pressure cylinder 
also produces another result. As the receiver pressure rises 
the work area of the high-pressure diagram is obviously de- 
creased, while that of the low-pressure diagram is increased. 
In a simple engine the area of the diagram becomes smaller 
the earlier the cut-off, and it should be noted that just the 
reverse of this occurs in the low-pressure cylinder of a com- 
pound engine. 

It is evident that the choice of the receiver pressure or 
of the point of cut-off in the low-pressure cylinder determines 
the relative areas of the high-pressure and low-pressure 
diagrams and it also determines the relative size of the two 
cylinders. The diagram of Fig. 93 shows that late cut-off 
in the low-pressure cylinder calls for a larger high-pressure 
cylinder than does early cut-off. 

The ratio of the piston displacement of the low-pressure 
cylinder to that of the high-pressure cylinder is called the 
cylinder ratio. Designating this ratio by R, and using other 
symbols as in Fig. 93, 

B=J|. ....... (61) 

The cylinder ratios chosen for real compound engines 
vary greatly in different designs and no given ratio has been 
proved the best for a given set of conditions. Normal 
practice gives the average values listed in Table VI, but 
cylinder ratios as high as 7 have been used with excellent 
results. 

TABLE VI 

Cylinder Ratios for Compound Engines 



Cylinder ratio 

Initial pressure (gauge) non-condensing . 
Initial pressure (gauge) condensing 



100 



3| 
120 
100 



120 



4* 

150 



COMPOUNDING 153 

The cylinder ratio to be used in a given case may be 
determined by any one of several considerations or by a 
combination of them, the latter being more often the case. 
Thus it may be deemed desirable to obtain the same amount 
of work from both cylinders; or to obtain equal temperature 
ranges; or to have cut-offs occur at the same fraction of 
the strokes; or to have the same total load on the two 
piston rods during admission; or to obtain the maximum 
possible uniformity of turning effort at the crank. The con- 
sideration of equal work is generally 
regarded as the most important. 

75. Indicator Diagrams and Mean 
Pressures. The idealized diagrams 
for a compound engine with clearance, 
with incomplete expansion in both 

cylinders, and without compression 

- - tp' i\a ■ rvu - i ■ Fig. 94. 

are given in Fig. 94. Ihe nominal 

total ratio of expansion would be L L -^-l H , but the total ratio 

of expansion taking account of clearance is 

Total ratio of expansion = L ni L , . . (62) 



and the cylinder ratio is 






The mean effective pressures can be found from each 
of the diagrams in the ordinary way and the indicated 
horse-power of each cylinder determined therefrom. The 
indicated horse-power of the engine is then equal to the sum of 
the values obtained for the separate cylinders. 

It is often convenient to refer the mean effective pres- 
sure of all cylinders to the low-pressure cylinder as though 
this were the only cylinder acting. In the simple form 
of diagram, such as that shown in Fig. 93, it is obvious 
that this could be obtained by measuring the area AbcDEA, 



154 STEAM POWER 

dividing by the length AE and multiplying by the scale of 
the spring, just as though the diagram were all produced 
in one cylinder with the piston displacement equal to V L . 
In the case of the diagrams given in Fig. 94 a similar method 
could be adopted, or the mean effective pressure of each 
cylinder could be determined separately and then the 
equivalent pressure which would give the same result on 
the low-pressure piston could be determined analytically. 

Assume for this purpose that the mean effective pressure 
of the high-pressure is equal to p H pounds per square inch, 
that the mean effective pressure of the low-pressure cylinder 
is equal to p L and that the cylinder ratio is R. The strokes 
of all cylinders of a multi-expansion engine are generally 
equal, so that the piston areas are in the same ratio as the 
cylinder volumes (piston displacements). In the case of 
a 2x engine, therefore, the area of the low-pressure piston 
is R times as great as that of the high-pressure piston, 
and the pressure required on the low-pressure piston to do 
the same work as that done by pressure pu on the high- 

pressure piston will be ^— . 
R 

In the case of a 2x engine therefore the total M.E.P. 

referred to the low-pressure cylinder is 

Vr^+Vl (64) 

This mean effective pressure acting on the low-pressure 
piston only would give the same indicated horse-power as 
is obtained with the two cylinders of the engine. 

In designing compound engines it is customary to 
determine the size of the low-pressure cylinder as though it 
were to do all the work expected of the engine by receiving 
steam at the highest pressure available and exhausting it 
at the lowest. The mean effective pressure which would 
thus be assumed to exist is the referred value p R just ex- 
plained. Having found the size of the low-pressure cylinder 



COMPOUNDING 155 

and the value of the referred M.E.P. the size of the high- 
pressure cylinder can be determined so that the work done 
by each cylinder will be just half of the total for which the 
engine is being designed. This size will have to be such 
that the high-pressure mean effective pressure referred to 
the low-pressure cylinder (i.e., p H +R) is equal to half the 
total mean effective pressure referred to that cylinder. 
That is, the size will have to be so chosen that 

f=f (65) 

ILLUSTRATIVE PROBLEM 

A double-acting compound engine is capable of developing 
500 I.h.p. The stroke is 18 ins.; revolutions per minute, 175; 
mean effective pressure referred to L.P. piston, 45 lbs. per square 
inch; cylinder ratio, 3^. Find cylinder diameters. 

From 

pLan 



I.h.p. 



33,000' 



500X33,000 
aLP - = 45Xl.5Xl75X2 =700(aPPrOX ' ); 



that 






4. 



— =30 ins. (approx.), 



with the cylinder ratio equal to 3j, 

«h.p. =-7tt =200 sq.ins., 
3.5 

d H .p. = \ i™ = 16 ins. (approx.). 

76. Combined Indicator Diagrams. When a compound 
engine is indicated, the diagrams of the two cylinders as 
drawn by the indicator are not directly comparable. The 
scales of pressure and volume are different on the two dia- 
grams, and correction must be made for this fact before the 



156 



STEAM POWER 



diagrams can be compared. It is customary to do this and 
to draw the average high-pressure and low-pressure diagrams 
on the same set of coordinates in order to determine how 
well they approximate the ideal diagram that would be 
obtained in one cylinder operating between the extreme 
limits of pressure. 

Diagrams approximating those that would be obtained 
from high- and low-pressure cylinders are shown at (h) 
and (I) respectively, in Fig. 95, and the result of drawing 




Fig. 95. 



both to the same scales is shown at the left of this figure. 
The curves xn and xi show the variations of quality along the 
two expansion curves. 

Drawing the two diagrams to the same scales in this 
way is known as combining the diagrams and the result is 
known as a combined diagram. 

The curves SS and S'S' added to the combined diagram 
are saturation curves. They do not, in general, form a 
continuous curve, because of the different quantities of 
steam contained in the two clearances and because any 



COMPOUNDING 



157 



moisture in the high-pressure exhaust is generally removed 
in the receiver. The volumes occupied by clearance steam 
at initial pressures are indicated by the points b f and B' 
respectively. The lengths b'S and B'S f approximately 
represent the volumes that would be occupied by cylinder 
feed when in each cylinder if dry and saturated. 

A combined diagram for a triple-expansion engine is 
shown in Fig. 96. The heavy lines give diagrams con- 
structed so as to represent 
as nearly as possible what 
may be expected to occur in 
the cylinders of such an en- 
gine, assuming perfect valve 
action and hyperbolic expan- 
sions and compressions. The 
dotted diagrams indicate the 
shapes that would be drawn 
by indicators applied to the 
real cylinders. The numer- 
ous sharp angles are due to 
overlapping of events, one 
cylinder suddenly starting to draw from a receiver while 
another is exhausting. It will be observed that the dotted 
diagrams do not contain any of these sharp angles, but 
that their general outline forms a fair average of them. 

The curve cd is a rectangular hyperbola drawn as a 
continuation of the assumed hyperbolic expansion line of 
the high-pressure cylinder. The failure of the expansion 
lines of the other cylinders to fall upon this curve is ex- 
plained by quality changes, different quantities of clearance 
steam in the different cylinders and withdrawal of moist- 
ure from steam exhausted to receiver before admission to 
the following cylinder, 




— .d 



Fig. 96. 



158 STEAM POWER 



PROBLEMS 



1. Find the size of the cylinders of a double-acting compound 
engine, which is to give 600 I.h.p., when using steam at a pressure 
of 150 lbs. per square inch absolute, and having a back pressure 
of 2 lbs. per square inch absolute. The cylinder ratio is to be 4, 
and the total ratio of expansion 12, piston speed 750 ft. per minute, 
and R.P.M. =150; diagram factor is 80%. 

2. Given a 200 H.P. compound Corliss engine with cut-off 
in the H.P. cylinder at 60% stroke. Ratio of expansion is 7; 
clearance is 7%; card factor is 70%; pressure at the H.P. cyl- 
inder is 165 lbs. absolute. Find 

(a) Cylinder ratio; 

(b) Theoretical and actual M.E.P.; 

(c) Determine size of four engines, and select the best one. 

l-f-% Q\ 

Note - /= % (c.o.)°+%ci Xcyl - ratio - 

3. Given a compound engine 18X40 ins., having a stroke of 
28 ins. Steam pressure is 165 lbs. per square inch absolute; 
cut-off in H.P. cylinder occurs at 62% stroke; clearance equals 
16%; back pressure equals 5 lbs. ; R.P.M. equal 150. Find 

(a) Cylinder ratio; 

(b) Ratio of expansion; 

(c) Actual M.E.P.; 

(d) I.h.p. 



CHAPTER X 
THE D-SLIDE VALVE 

77. Description and Method of Operation. The simple 
D-slide valve, shown in place in Fig. 97, is so named because 
of the similarity of its section to the letter D. It is located 
in the steam chest, rides back and forth upon its seat and 




Fig. 97. 



serves to connect the two ports alternately with steam 
and exhaust spaces respectively in order to give the neces- 
sary distribution of steam. 

The valve has to perform the following functions for 
each end of the cylinder during each revolution of the 
engine : 

(1) It connects the proper port to the steam space or 

159 



160 STEAM POWER 

steam chest at such a time that steam can enter the cylinder 
as the piston moves away from the head. 

(2) It shuts off this port and thus cuts off the supply 
of steam when the piston has completed a certain definite 
fraction of the stroke. 

(3) It connects the port with the exhaust cavity shortly 
before the piston reaches the end of the stroke, thus effecting 
" exhaust " or " release "; and 

(4) It shuts off the port again when the piston has com- 
pleted the proper fraction of the next stroke, thus trapping 
in the cylinder the steam which is compressed during 
the remainder of the stroke. 

Engine Crank rl 
Main Co nnecting 




Fig. 98. 

It is obvious that the valve must be reciprocated upon 
its seat and that its motion must be connected with that of 
the piston in some way so that the proper phase relation 
may be retained. This could be effected by the system 
shown diagrammatically in Fig. 98, a small crank operating 
on the end of a connecting rod giving the valve its short 
stroke just as the main crank fixes the longer stroke of 
the piston. Such an arrangement would, however, be 
very inconvenient with many real engines, as the valve would 
be located too far from the center line of the cylinder. 

It is customary to use what is knbwn as an eccentric 
for the purpose of operating the slide valve. The parts 
and arrangement of an eccentric, together with an illus- 
tration of the way in which it is mounted on the shaft of 



Strap 




Fig. 99.— Parts of Eccentric. 



161 



162 



STEAM POWER 




THE D-SLIDE VALVE 



163 



L 



an engine are shown in Figs. 99, 100 and 101. The motion 
it gives the valve is exactly the same as that imparted 
by the crank first assumed, and it can 
easily be shown that it is the exact 
equivalent of such a crank. 

Assume, for example, a crank such 
as that shown in Fig. 98 with a length 
of arm or throw equal to a. If the 
crank pin is made larger while other 
parts of the crank remain the same, as 
shown in Fig. 102, the crank mech- 
anism is not essentially altered; the mo- 
tion which it would impart to a connecting rod is not 
changed. If this process of enlarging the pin be continued 




Fig. 101. — Eccentric 
on Vertical Engine. 




(<0 (d) 

1 
Fig. 102. — Equivalence of Crank and Eccentric. 

until the pin has become large enough to surround the 
shaft and if the crank arm be then removed so that what 
was the crank pin is fastened directly on the shaft, an 



164 



STEAM POWER 





^sssssssss^s^s 




Fig. 104. — Slide Valve without Lap, 



Live Steam Space. 

Exhaust Cavity Connected 
to Exhaust J?ipe 




Fig. 105. 



THE D-SLIDE VALVE 165 

eccentric results. It is the exact equivalent of the original 
crank; its center, which is the center of the crank pin, 
revolves about the center line of the shaft in a circle with 
a radius a just as in the original mechanism. 

The eccentric makes it possible to place a short crank 
(short arm) upon a large diameter shaft without having to 
cut the shaft away as shown in Fig. 103, and it is therefore 
very useful for driving valves. 

78. Steam Lap. The simplest possible form of D-slide 
valve would just reach the outer edges of the ports when 
in its central position as shown in Fig. 104. The crank 
driving it (that is the crank equivalent to the eccentric 
which would probably be used in a real case) would have 
to be located 90° ahead of the engine crank in the direction 
of rotation, as can easily be seen by consulting Fig. 105, 
which illustrates the mechanism in various critical positions. 
The illustration shows that such a valve would give full 
stroke admission, thus producing a rectangular cycle which 
has already been shown to be very inefficient as a means of 
obtaining work from the heat used in 

i> • , Steam-Space 

IOrmmg Steam. ^Valve Travel? 

If cut-off is to occur before the - ^^P^T^"-^— ^ - 
end of the stroke, the edge of the ^^ JXdm 
valve must return and close the port cSy* 
before the piston reaches the end of its 
stroke. But since the crank mechan- - 



Piston Travel 



ism does not permit the valve to Fig. 106. 

remain stationary in any one posi- 
tion, such early cut-off could only occur if the valve 
over-traveled, as shown in Fig. 106, and this would un- 
fortunately result in connecting the working end of the 
cylinder to exhaust and in admitting steam to the other 
side of the piston at such a time as to oppose the piston's 
motion. The solution of the difficulty lies in making 
the valve longer, so that when in its central position it 
overlaps the outer edges of the ports as shown in Fig.. 107.. 



166 



STEAM POWEE 



The amount of overlap of the outer edge is called the out- 
side lap, and when steam is admitted by the outer edges of 
the valve, as in the case under discussion, it is also called 
the steam lap. 

With such an arrangement the valve must be drawn 
out of its central position by the amount of the lap when 
the piston is at the end of its stroke as shown by a in Fig. 





Fig. 107.— Steam and 
Exhaust Lap. 



Fig. 108. — Lap a and Lap 
Angle a. 



108 in order that steam may be admitted just as the 
piston starts to move. It follows that the crank driving 
the valve must be more than 90° ahead of the engine 
crank and that it must be ahead by the angle required 
to move the vaive a distance equal to the outside lap. 
This angle, represented in the figure by a, is called the lap 
angle. 

79. Lead. In real engines it is further desirable to start 
the admission of steam just before the piston arrives at 
the end of its stroke. This assists in bringing the moving 
parts to rest, raises the pressure in the clearance to full 
value before the piston starts, and gives a wider opening 
through which the steam can flow during the early part of 
the stroke, thus reducing wiredrawing and loss of area at 
the top of the diagram. If the valve is to open before the 
piston reaches the end of its stroke, the crank driving it 
must be shifted still further ahead of the engine crank. 
It must be shifted ahead by an angle which will draw 
the valve through the distance which will give the desired 
opening of valve with the piston at the end of its stroke 
as shown by b in Fig. 109. The angle required, indicated 



THE D-SLIDE VALVE 167 

by j8, is known as the angle of lead, and the width 
of the steam opening with engine crank on dead center, 
i.e., the distance b, is known as the lead. The lead 
varies from less than t^ in. on small engines and with low 
speeds up to over J in. on large engines and with very high 
speeds. 

80. Angle of Advance. The eccentric or valve-operating 
crank must be ahead of the engine crank by an angle equal 
to 90°+ angle of lap a + angle of lead j8, as can be seen 




Fig. 109. — Lead b and Lead Angle /?. 

by an inspection of Fig. 109. The sum of a and is called 
the angle of advance and will be represented by 8. This 
is the number of degrees in excess of 90 by which the eccen- 
tric leads the engine crank. 

Fig. 109 shows that cut-off in an engine fitted with 
a valve having lap and lead must occur when the engine 
crank has turned through an angle equal to 180 — 2a, because 
the valve will then have returned to the closed position. 
Apparently, cut-off can be made to occur at any point 
in the stroke by properly choosing the value of a, but it 
will be discovered later that the exhaust events set a limit 
to increase in the value of this angle and hence do not per- 
mit of cut-off occurring earlier than a certain fraction of the 
stroke. 

81. Exhaust Lap. Inspection of Fig. 105 will show that 
the simple valve without lap originally assumed will give 
no compression, because the cylinder end is connected to 
the exhaust cavity for the entire stroke. Inspection of all 



168 STEAM POWER 

the changes which have been suggested in the subsequent 
paragraphs will show further that if the inner edges of the 
valve are left in the original positions the exhaust events 
will be considerably distorted in the case of a valve having 
steam lap and lead. 

This trouble may be remedied by moving the inner 
edges of the valve closer together, making the exhaust 
cavity in the valve shorter and giving inside lap as shown 
in Fig. 107 by b. When the inner edges of the valve control 
exhaust, as in the case of the valve under discussion, this 
inside lap is also called exhaust lap. 

The length of the valve, the lap and the lead are gen- 
erally chosen so as to give the desired arrangement of 
admission and cut-off and then the exhaust edges are so 
located as to give desirable release and 
compression. In some forms this necessi- 
tates the use of an exhaust cavity in the 
Fig. 110. valve such as that shown in Fig. 110. 

The amount by which the edges of the 
valve fail to meet the inner edges of the port is spoken of as 
negative inside lap. This dimension is indicated by c in the 
figure. 

It should be noted particularly that all measurements 
of lap are made with the valve central on its seat and 
that the measurement of lead is made with the piston at 
the end of its stroke, i.e., with the engine crank on dead 
center. 

82. The Bilgram Diagram. The action of all slide 
valves could be studied by means of drawings of the actual 
mechanism, as has been done in preceding paragraphs, but 
such a method is time and space consuming. Numerous 
diagrams such as the Elliptical, the Sweet, the Zeuner and 
the Bilgram have been developed for the purpose of simpli- 
fying and expediting such studies and, when properly 
understood, they are very convenient. The scope of this 
book does not permit a discussion of all of these diagrams 



THE D-SLIDE VALVE 



169 



and, since the Bilgram diagram is probably the most gener- 
ally applicable, attention will be confined to it. 

The construction of this diagram is illustrated in Fig. 
111. The point represents the center of the engine 
crank shaft and the two circles drawn about this point 
as a center represent respectively the paths traveled by the 




Fig. 111. 



pin of the valve crank and the pin of the engine crank. 
These circles are drawn to any convenient scales. 

The diagram is conventionally drawn in such a way 
that the line OM represents the head end dead center 
position of the crank and in all subsequent paragraphs the 
relative positions shown by the small sketch in Fig. Ill 
will be assumed. The cylinder will be assumed to the 



170 STEAM POWER 

left of the shaft and the engine will be assumed to run 
" over." 

With the crank in position OM , the eccentric (equivalent 
crank) must be in the position OB, ahead of the crank by 
an angle 90°+«+j3 = 90+<5. The valve must then be 
displaced to the right of its central position by an amount 
represented by the distance DB, if a small correction for 
" angularity " of the valve connecting rod be neglected. 
As rotation continues, horizontal distances corresponding 
to this line will always give the instantaneous valve dis- 
placements. For position OB' , for instance, the valve dis- 
placement will be D'B' '. 

If the angle 8 is now laid off above OX, locating the point 
Q as shown, a perpendicular QE dropped upon OX from this 
point will equal in length the line DB, and will therefore 
show the valve displacement when the crank is in head end 
dead center position OM. This must be true, because the 
triangles QOE and BOD are similar and have the sides 
OQ and OB equal to the radius of the same circle. 

The perpendicular QE is really a perpendicular dropped 
upon the extension of the line representing the crank posi- 
tion, and it is a general property of this diagram that a line 
starting at Q and perpendicular to the line representing 
any chosen crank position (or an extension of that line) 
will show by its length the displacement of the valve when 
the crank is in the chosen position. Thus assume the engine 
crank to rotate through the angle 7 to the position OM'. 
The eccentric will have rotated to B' and the valve dis- 
placement will be represented by D'B' . A perpendicular 
drawn from Q upon OX', the extension of the crank posi- 
tion, gives QE' equal to B'D' and hence representing the 
valve displacement to the same scale. 

This construction drawn for different crank positions 
OA, OM, OM x , OM 2 , etc., is shown in Fig. 112, the dash- 
dot radial lines about Q representing the various values of 
the valve displacement. The number on each of thesf. 



THE D-SLIDE VALVE 



171 



lines indicates the crank position to which it corresponds. 
It will be seen that the displacement increases in value 
until the crank position OM3 is reached, after which it 
decreases again. 



Steam Lap 
Circle- 



Steam Lead 




Fig. 112. 



Since the opening 
ment minus the lap, as 
by which the valve is 
found by subtracting 
placement the amount 
head end dead-center 
to lap plus lead, and is 



to steam is equal to the displace- 
shown in Fig. 109, the actual amount 
open for any crank position can be 
from the corresponding valve dis- 
of lap possessed by the valve. For 
position, the displacement is equal 
shown by QE in Fig. 112. Subtract* 



172 STEAM POWER 

ing the lead EF, the remainder FQ gives the lap of the valve. 
A circle drawn about Q with radius equal to QF (or a circle 
drawn about Q and tangent to the line L) will cut off of 
the lines representing valve displacement the amount 
representing the part of each displacement used in over- 
running the lap of the valve. The remainders, that is the 
parts of the lines radiating from Q in Fig. 112 which are 
outside of the lap circle, must then represent the amounts 
by which the valve port is actually open. 

It will be observed that the valve is open by the amount 
of the lead when the crank is on dead center, position OM. 
The crank position for which the valve displacement is just 
equal to the lap, and hence at which the valve is just begin- 
ning to open, can be found by drawing a tangent through 
to the lower side of the lap circle and then extending 
it to give the crank position OA in Fig. 112. 

As the crank rotates clockwise from this position, the 
valve opens wider until, when position OM3 is reached, 
the greatest valve opening exists. Further rotation results 
in partial closure of the valve and, when the crank has 
finally rotated into position OC, the valve has just closed, 
that is, cut-off has occurred, the displacement being just 
equal to QG, the steam lap. 

Thus this diagram, as so far developed, indicates crank 
positions for admission and cut-off and the values of valve 
displacement and valve openings for all intermediate 
crank positions. 

ILLUSTRATIVE PROBLEM 

A certain valve has an external steam lap equal to 1J ins. 
The lead is ^ in. and the throw of the eccentric is 2\ ins. (a) Con- 
struct such parts of the Bilgram diagram as are necessary to 
indicate ^head end" crank positions for admission, maximum 
valve opening and cut-off. (b) Indicate on this diagram the 
amount of valve opening at various crank positions between 
admission and cut-off. (c) Determine the value of the angle of 
advance. 



THE B-SL1DE VALVE 



173 



Draw a circle with radius equal to the eccentric throw, 2\ 
ins., using any convenient scale. This circle is designated by abed 
in Fig. 113. Draw about the same center another circle of any 
convenient size. Draw in the horizontal diameter ac and extend 
as shown. On the right-hand side of the circle draw the line ef, 



Steam Lap 
Circle 




Fig. 113. 



parallel to the horizontal axis and a distance above it equal to the 
lead, Y6 i n -, to the same scale as that chosen for eccentric circle. 
The steam lap circle must have its center Q on the upper right- 
hand quadrant of the eccentric circle, and it must be tangent 
to the line ef. Its radius must equal the steam lap, If in. to scale. 
Therefore, with compass points set the proper distance apart, find 
the center Q, about which a lf-in. radius circle will just be tangent 
to the line ef, and draw the steam lap circle. 



174 



STEAM POWER 



The crank position at admission is found by drawing the line 
AO so that, if extended, it is tangent to the lower side of the 
steam lap circle. 

The crank position at cut-off is found by drawing the line 



H.E. Adm 




I.E. Compression 




Fig. 114. 



M""0 in such position that it is tangent to the upper part of 
the steam lap circle. 

The crank position for maximum valve opening is found by 
drawing the line M"0 in such position that a line through QO 
will be perpendicular to it. The amount of valve opening at this 



THE D-SLIDE VALVE 175 

crank position is shown by the length of the part of this per- 
pendicular line outside of the steam lap circle, i.e., the distance 
Og interpreted according to the scale chosen for eccentric and 
steam lap circles. 

When the crank is in position M'O, the length of hi, interpreted 
to scale, gives the amount by which the valve is open to steam. 

When the crank is in position M'"0, the length of jk, inter- 
preted to scale, gives the amount by which the valve is open to steam. 

The angle indicated by 5 is equal to the angle of advance 
because of the property upon which the construction of this 
diagram is based. 

83. Exhaust and Compression. The exhaust edge events 
can be shown on the Bilgram diagram by a method similar 
to that used for the steam edge events. The direction in 
which valve displacements occur are indicated in the upper 
part of Fig. 114 in which the crank and eccentric circles 
have been drawn to such scales that they coincide. In- 
spection of the small sketch in the lower part of the figure 
will show that head end release must occur when the valve 
has traveled a distance equal to the inside lap to the left 
of its central position. A crank position OR drawn tangent 
to the lower part of a circle about Q with radius equal to 
the inside lap will, therefore, be the crank position at re- 
lease. Clockwise rotation from this position will result 
in a wider opening to exhaust until position OM\ is reached, 
after which the valve will begin to close. Final closure 
will occur when the crank reaches position OK, the exten- 
sion of which is tangent to the top of the exhaust lap circle. 
At that time the valve will have returned (moving from left 
to right) and will still have to move a distance equal to 
the exhaust lap before attaining a central position. 

ILLUSTRATIVE PROBLEM 

Given the exhaust lap of a D-slide valve equal to f in.; the 
steam lap \\ ins.; the throw of the eccentric, 2 ins.; and the 
lead \ in. Find the angle of advance, the maximum port opening 
to steam and to exhaust, and the crank positions of cut-off, release, 
compression and admission for the head-end of the cylinder. 



176 



STEAM POWER 



Draw the eccentric (and crank) circle with a radius equal 
to 2 ins., and draw the horizontal diameter as in Fig. 115. 

Draw a horizontal line in the upper right-hand quadrant at a 
distance of |+1J ins. above the horizontal diameter. Locate 
the point Q at intersection. 




Fig. 115. 



Draw the steam lap circle with a radius 1^ in. and the exhaust 
lap circle with a radius f in. 

The angle of advance is the angle between OQ and the hori- 
zontal. 

The maximum opening to steam is given by the distance 
Oa=f in. The maximum opening to exhaust is given by the 
distance 06 = If in. 

The crank positions shown are obtained by drawing lines 



THE D-SLIDE VALVE 



177 



tangent to the lap circles. A represents admission; C, cut-off; 
R, release, and K, beginning of compression. 

The piston positions at the times of these events are given 
to reduced scale by vertical projection. 

84. Diagram for Both Cylinder Ends. The complete 
diagram for the head end of cylinder is shown in Fig. 114 with 
all critical crank positions marked. The positions for the 
crank end of the cylinder can be found in a similar way by 
constructing a diagram in which the point Q and the lap 
circles are located in the opposite quadrant. The resulting 




Fig. 116. 



diagram for both cylinder ends, with laps the same for both 
ends of the valve, is given in Fig. 116. 

85. Piston Positions. The valve events might be studied 
entirely in conjunction with crank-pin positions, but it is 
more convenient and customary to consider them in connec- 
tion with piston positions. Piston positions corresponding 
to different crank-pin positions could be found by drawing 
the mechanism to scale for each different position as shown 
in Fig. 117 for piston positions 1 and 2. 

It is obvious that this would involve a great deal of work 
and that, if drawn to large scale, it would consume a great 



178 



STEAM POWER 




! \ 



8 



© 



e 



deal of space. Further, it is 
convenient to be able to locate 
relative piston positions on the 
line which serves as the hori- 
zontal diameter of the crank 
circle of the Bilgram diagram. 

The method used depends 
upon the fact that the motion 
of the crosshead is exactly the 
same as that of the piston, 
so that if the motion of the 
crosshead end of the connect- 
ing rod can be followed, it 
will be equivalent to following 
the motion of the piston itself. 
It should also be noted that 
the diameter of the crank cir- 
cle must be equal to the 
stroke of the engine. 

Assume now, that the point 
b in Fig. 117 be taken to rep- 
resent the position of the pis- 
ton when it is really in posi- 
tion 1. When the piston has 
moved to position 2, the cross- 
head will have moved from a 
to a' and the crank pin from 
b to b' . If with a' as a center 
the connecting rod be swung 
down to the horizontal its 
right-hand end will arrive at 
the point c. The distance be 
must then represent the dis- 
tance that crosshead (and pis- 
ton) have moved from dead- 
center position because ab and 



THE D-SLIDE VALVE 



179 



a'c both represent the length of the connecting rod and c 
must therefore be as far to the right of b as a' is to the right 
of a. The point c may therefore be taken to represent 
pistor' position when the connecting rod is in the position 
a'b'. 

In general, if the horizontal diameter of the crank 
circle be taken to represent the stroke of the engine, the pis- 
ton position corresponding to any crank position can be 
found by taking a radius equal to the connecting-rod length 
(to the same scale as the circle) and striking an arc from the 




Fig. 118. 



crank-pin position, using a center on the horizontal line on 
the cylinder side of the crank circle. 

An approximate method is also used for finding the piston 
position. Instead of projecting down from the crank-pin 
position with an arc, such as Vc in Fig. 117, a vertical line 
through the crank-pin position is used. Such a line would 
give c' as the piston position when c is really correct. This 
method would give accurate results with a connecting rod 
of infinite length. For ordinary lengths of rod, however, 
the results are far from correct. The error is said to be due 
to the angularity of the connecting rod. 

The effect of the angularity of the connecting rod is 
shown in Fig. 118 for different positions. On the outstroke 
the piston is always farther ahead than the rectilinear pro- 



180 



STEAM POWER 



jection would indicate. On the return stroke the piston is 
always behind the position indicated by rectilinear projection. 




Fig. 119. 



86. Indicator Diagram from Bilgram Diagram. Since 
the piston positions corresponding to different crank posi- 



THE D-SLIDE VALVE 181 

tions can be determined, it is a comparatively simple matter 
to construct the indicator diagram which theoretically 
would be given by an engine fitted with a valve of certain 
dimensions. It is necessary to assume the upper and lower 
pressure and also to assume the form of the expansion 
and compression curves. These are generally taken as 
rectangular hyperbolas. 

The method of constructing an indicator diagram from 
the Bilgram diagram is shown in Fig. 119. The crank- 
pin positions for admission (A), cut-off (C), release (R) 
and beginning of compression (K) are first found. These 
pin positions are then projected to the horizontal diameter 
by means of arcs with radius equal to the connecting-rod 
length and with centers on the line MM produced to the 
left. The intersections a, c, r and k indicate the piston 
positions at which the corresponding events occur. These 
are then projected vertically downward to intersect the 
proper pressure lines and the card is drawn through the 
intersections. 

Diagrams constructed in the same way, but for both 
head and crank ends, are given in Fig. 120. A symmetrical 
valve was assumed, that is, one built exactly alike on head 
and crank ends. The diagrams show that such a valve 
cannot give the same results for both cylinder ends because 
of the effect of the angularity of the connecting rod. It 
is most evident in the case of cut-off. The cut-off in this 
case occurs just before three-quarter stroke for the head end 
and just after half stroke for the crank end of the cylinder. 
All other events are distorted in the same way, but the actual 
lengths of the variations are not as great as in the case of 
the cut-offs and therefore the distortion is not as obvious. 

The effect of the angularity of the connecting rod upon 
the diagrams can be remembered easily if it is noted that all 
valve events occur later with respect to piston position on 
the outstroke and earlier on the instroke than they would 
with a connecting rod of infinite length. 



182 



STEAM POWER 



It is possible to " equalize " the cut-offs, that is, make 
them occur at the same fraction of the stroke by using 
unequal steam laps at opposite ends of the valve, but this 
will result in still further distortion of admissions, as can be 
seen by constructing a Bilgram diagram for this case. 
Similarly, the compressions can be equalized by the use of 



•pression 



H.E. Admission 




C.E. Admission 



Fig. 120. 



unequal exhaust laps, but this results in distortion of the 
release events. 

Various linkages have been developed which are so 
arranged that they distort the motion of the valve to just 
the extent necessary to counterbalance the effects of the 
angularity of the connecting rod. The scope of this book 
does not, however, permit a discussion of such valve 
gears. 



THE B-SLIDE VALVE 183 

87. Limitations of the D-slide Valve. The simple 
valve discussed in the preceding paragraphs has numerous 
limitations and is therefore only used on small and cheap 
engines, or in cases where economy in the use of steam is 
not essential. This valve, when used with steam entering 
over the outside edges as previously considered, is pressed 
to its seat by the live steam acting over its entire upper 
surface. This pressure is practically unbalanced, as the 
greater part of the lower surface of the valve is subjected 
to the low pressure of the steam being exhausted. As a 
result the friction to be overcome in moving the valve is 
very great and there is an appreciable loss from this source. 

Further, the shape of the valve makes necessary the use 
of long ports which form part of the cylinder clearance 
and which are alternately exposed to live and to exhaust 
steam with results previously discussed. These ports can 
be decreased in length by increasing the length of the valve, 
but this in turn increases the area exposed to high pressure 
and hence increases the friction loss. 

It can be shown by means of the Bilgram diagram 
that, if a cut-off earlier than about f stroke is desired, 
the angle of advance, the amount of steam lap and the size 
of the eccentric must all be made very great. This results 
not only in large friction losses, but also in very early release 
and compression, because of the great angle of advance. 
As a result, slide valves of the simple D type are seldom used 
when a cut-off earlier than \ to \ stroke is desired. It 
should be remembered in this connection that the simple 
engine generally gives its best economy with a cut-off of 
about J stroke. 

The drawing of lines representing the opening of the 
valve to steam as in Fig. 112 will show that this simple 
valve is further handicapped by the very slow opening 
and closing of the steam ports, causing a great amount of 
wire drawing with a corresponding loss of diagram area. 
In order to get an adequate opening to steam the valve 



184 



STEAM POWER 



.Steam Spaces, 

.Exhaust i 



must also be given a great displacement and, since this 

occurs under great pressure, it results in great friction loss. 

The unbalanced feature can practically be overcome 

by rolling up the valve 
and ports about an axis 
parallel to the length of 
the cylinder. This gives 
what is known as a pis- 
ton valve, shown dia- 
grammatically in Fig. 
121 

It can also be par- 
tially overcome by using 
a balance plate or ring of some kind between the top of 
the valve and the inside of the steam-chest cover, so 
arranged that live steam is excluded from the greater part 
of the upper surface of the valve. Valves of this type 
are generally called balanced slide valves and are used on 
many high- and medium-speed engines. 

The valve travel required for obtaining a given opening 




■Steam Ports 



Fig. 121— Piston Valve. 




Fig. 122.— Allen Double Ported Valve. 



can be decreased and the rate of opening and closing can be 
increased by the use of multiported constructions. These 
are so arranged that two or more ports open or close at the 
same time, so that the total movement required for a given 
opening is divided by the number of ports and the rate of 
opening and closing is multiplied in the same proportion. 
One simple type of double-ported valve is illustrated in 
Fig. 122. 

When several ports are used the valve often becomes 



THE D-SLIDE VALVE 185 

a rectangular frame crossed by a number of bars and is 
known as a gridiron valve, because of its appearance. Such 
valves are often combined with balance plates and give 
very satisfactory results. 

A number of designs of slide valves have been developed 
for the purpose of making cut-off independent of the other 
events. Many of these use a separate cut-off valve which 
either controls the steam supply to the main valve or else 
rides on the main valve and controls cut-off by covering 
ports in that valve. Devices of the latter type are called 
riding cut-off valves. They are either driven by separate 
eccentrics, or by linkage from the eccentric controlling 
the main valve, the linkage being so arranged as to give 
the proper relative motion between main and auxiliary 
valves. In such designs the main valve is proportioned 
so as to give the desired admission, release and compres- 
sion and the cut-off is then taken care of by proper adjust- 
ment of the cut-off valve, 

88. Reversing Engines. It was shown in one of the early 
paragraphs of this chapter that the eccentric must be set 
90°+ angle of advance ahead of the crank, ahead meaning 
in the direction of rotation. To cause the engine to revolve 
in the opposite direction, that is, to " reverse " the engine, 
it is therefore only necessary to shift the relative positions 
of eccentric and crank so that the eccentric leads the crank 
by 90°+ 8 in the new direction of rotation. This corre- 
sponds to shifting ahead (in first direction of rotation) 
through an angle equal to 180 — 25 or shifting backward 
through an angle equal to 180+25, as can be seen by inspec- 
tion of Fig. 109. 

In practice it is generally more convenient to use two 
eccentrics, one set properly for rotation in one direction 
and the other set properly for rotation in the opposite direc- 
tion. This arrangement is shown diagrammatically in 
Fig. 123. This figure is drawn for a vertical engine and in 
such position that the engine is on crank-end dead center. 



186 



STEAM POWER 




Reverse or 
Weight Shaft 

Valve Stem 



The point P represents the position of the center of the crank 

pin; the point / represents the position of the equivalent 
crank (center of eccentric) which 

/ / ^v \ drives the valve for " forward," 

" ahead " or clockwise rotation; and 
the point b represents the position of 
the equivalent crank which drives the 
valve for " backing/' " reverse," or 
counter-clockwise rotation. 

The real mechanism, in one of its 
numerous forms known as the Stephen- 
son Link Gear, is shown in perspective in Fig. 124. The 

forward eccentric corresponds to / of Fig. 123 and the 

backing eccentric corresponds 

to b of that figure. The 

eccentric rods are fastened 

to opposite ends of a curved 

" link " and move the valve 

through a " link block " 

fastened to the end of the 

valve stem. In the position 

shown in the figure the link 

is in such position that the 

forward eccentric operates 

practically directly on the 

valve stem so that the valve 

motion is practically entirely 

governed by that eccentric. 

If the reverse shaft were to 

be rotated clockwise into the 

backing position, the " sus- 
pension rods " would pull 

the link over until the eccen- 




Clockwise 
Rotation 



Fig. 124. — Stephenson Link Gear. 

trie rod of the backing eccentric was directly under the 
valve stem. Under such conditions the valve motion 
would be controlled almost entirely by the backing 



THE D-SLIDE VALVE 187 

eccentric and the engine shaft would rotate counter-clock- 
wise. 

If the mechanism were so set that the link block occupied 
a position on the link between the ends of the two eccentric 
rods, the valve motion would be controlled by both eccentrics 
and would be a compromise between the motions given by 
either eccentric separately. It is characteristic of this gear 
that the cut-off is latest when either one or the other eccentric 
is fully " in gear " and that it becomes earlier as the link 
block approaches the center of the link. With the link 
block in the center of the link the valve does not open at 
all, i.e., the cut-off occurs at zero stroke. 

There are numerous other forms of link gears, the best 
known being the Gooch, the Allan and the Porter- Allen. 
There are also numerous reversing mechanisms known as 
radial gears in which the motion of the valve is controlled 
by means of a " radius rod " which can be set to give the 
desired valve motion. • The valve motion is obtained in- 
directly through the radius rod from an eccentric, from the 
crank, or from the connecting rod. The limits of this 
book do not permit a detailed discussion of these forms. 

89. Valve Setting. From what has preceded it will 
be evident that it is not only necessary that a valve and 
its seat and driving mechanism be correctly designed, but 
also that the various parts must be correctly connected up 
in order that the valve may move in its proper phase rela- 
tion with respect to the piston. 

Adjusting the mechanism in such a way that the proper 
phase relations are obtained is known as setting the valve. 
This can be done with fair accuracy by a simple study of 
the mechanism in various positions, as will be shown below, 
but it is always advisable to check the setting by means of 
indicator diagrams taken after the setting is completed. 
Such diagrams will often show errors of such character or 
size that they cannot be determined by measurement on 
an engine which is not operating. 



188 



STEAM POWER 



Before beginning operations it is always advisable to go 
over the entire engine carefully and to eliminate excessive 
lost motion at all pins and bearings in order that the relative 
positions of parts obtained while setting the valve may 
approximate those which will be obtained when the engine 
is in operation. The effect of lost motion will be appreciated 
after a study of Fig. 125. Assume that all parts of the 
mechanism are tight except the crank-pin end of the con- 




Fig. 125. 



necting rod as shown. If, for instance, the engine is rotated 
by hand by turning the fly-wheel, the crank will pull the 
piston mechanism and the piston will be drawn into the 
position shown in the upper half of the figure when the crank 
has turned through an angle a. On the other hand, when 
the engine is operating under steam, the piston will push 
the crank pin around and will occupy a position such as 
that shown in the lower half of the figure when the crank 
has been turned through the same angle a. Obviously, 
the piston can occupy two very different positions for the 
same crank position, and a valve setting based upon the 
conditions shown in the upper part of the figure might be 



THE D-SLIDE VALVE 189 

very incorrect when used under the conditions shown in 
the lower part of the figure. 

Lost motion in any part of the mechanism can produce 
analogous results and it is therefore necessary to remove as 
much of it as possible before attempting to set the valve. 
It is practically impossible to eliminate all lost motion, as 
there must be sufficient clearance at all bearing surfaces 
to accommodate a film of oil, and this alone would make 
necessary the taking of indicator diagrams for the check- 
ing of valve settings, even if it were possible to set perfectly 
by measurement for stationary conditions. 

In general, there are two adjustments which can be made 
in setting a plain slide valve. The length of the valve 
stem or eccentric rod can be changed and the eccentric 
can be shifted around the shaft. It is necessary to under- 
stand the effects of each of these adjustments. 

Changing the length of the valve stem is equivalent to 
shifting the valve upon its seat without 
moving the engine as shown in Fig. 
126. In this figure the valve is shown 
in its central position by full lines. The 
lap is the same at both ends. If, now, Fig. 126. 

the valve is worked to the right upon 
its stem by adjustment of the nuts shown, until it reaches 
the dotted position, the head-end lap will have been de- 
creased and the crank-end lap will have been increased by 
the same amount. This would make admission earlier and 
cut-off later for the head end and admission later and cut- 
off earlier for the crank end. Obviously, the effects of 
changing the length of the valve stem are opposite for the 
two ends of the cylinder. 

Shifting the eccentric about the shaft simply changes 
the time relation between valve motion and piston motion; 
it does not alter the valve motion itself. If difficulty is 
experienced in realizing the truth of this statement, it is 
only necessary to draw several Bilgram diagrams for the 




190 STEAM POWER 

same valve, but with different angles of advance, and then 
to construct indicator diagrams for both cylinder ends in 
every case. It will be discovered that shifting the eccentric 
ahead in the direction of rotation, for instance, will make 
all events occur earlier with respect to piston position for 
both ends of the cylinder. 

In setting a plain slide valve which is built symmetrical 
about a central axis, i.e., same inside and outside lap at 
each end, it is first necessary to adjust the length of the valve 
stem. This may be done by removing the steam-chest 
cover so as to expose the valve and then rotating the engine 
slowly by hand and observing the distance traveled by the 
valve on each side of its central position. This is con- 
veniently done by observing the distance between the outer 
edge of the steam port and the outer edge of the valve when 
the valve is fully open at each end. If the valve travels 
further toward the head end than it does toward the crank 
end, with reference to the port edges, the valve stem must 
be shortened; if it travels further toward the crank end 
the stem must be lengthened. 

In making these adjustments it is advisable to turn the 
engine only in the direction in which it is going to rotate, so 
that any lost motion in the valve mechanism will have 
approximately the same effect as when the engine is opera- 
ting. 

When the length of the valve stem is correctly adjusted, 
the eccentric must be so set on the shaft as to give the proper 
angle of advance. This is commonly done by shifting it 
about the shaft until the proper value for the steam lead has 
been obtained. In order to determine the value of the lead 
it is necessary to be able to set the engine on each dead center. 
This can be done approximately by turning the engine until 
the crosshead has come to either end of its stroke, but it 
will be found by trial that the fly-wheel and shaft can be 
turned through a very large angle at each end of the stroke 
without causing an appreciable motion of the crosshead, 



THE D-SLIDE VALVE 



191 



so that this method is not very satisfactory for the purpose 
of adjusting the eccentric. It is customary, therefore, 
to work in such a way as to give a more accurate determina- 
tion of shaft and crank positions for dead center. 

The engine is rotated until the crosshead has been 
brought near one end of its stroke, as shown in Fig. 127, 
and a mark is then scribed across the crosshead and guide 
as at ab. An arc xy is then marked on the fly-wheel by 
means of a tram such as that shown, the end c being placed 




*Mmti& 



K-Trai 



Fig. 127 



at point P on some solid part of foundation or floor. The 
engine is then rotated, clockwise in the figure, until the 
crosshead has reached the end of its stroke and returned 
to such a point that the marks on crosshead and guides 
again coincide, as shown by dotted positions in the figure. 
The arc x'y' is then scribed on the fly-wheel with the tram, 
the end c again bearing on the point P. A point z is then 
found by bisecting the arc ef and when this point is brought 
under point d of the tram the crank will obviously be at 
crank-end dead center and the piston at the crank end 



192 



STEAM POWER 




(a) Perfect Cards for Slide Valve Type. 




(b) Actual Card; Small Engine. Center Line of Valve on Center 
Line of Seat; Eccentric Advanced to Give Normal Lead of 
0.05 inch. Engine Running Over. 




(c) Same Setting as (6) except Engine Running Under. 
Fig. 128. 



THE D-SLIDE VALVE 



193 




(d) Angular Advaice of Eccentric Increased. Valve Stem Length 
Same as in 6) and (c). Lead 0.375 Inch, 




(e) Angular Advance of Eccentric Decreased so as to Give Negative 
Lead of 0.5 Inch. Length of Valve Stem Unchanged. 




(/) Length of Valve Stem Changed; Angle of Advance as in (6). 

Fig. 128. 



194 STEAM POWER 

of its stroke. A point on the fly-wheel diametrically opposite 
to z is next found, so that when it is brought under point 
d of the tram the engine will be on head-end dead center. 

It is probable that more accurate results are obtained 
by rotating the engine in a direction opposite to that in 
which it rotates under steam, because lost motion is then 
taken up in the same direction as when working, but when 
the whole process of valve-setting is considered it is ques- 
tionable whether this is the correct direction of rotation. 
Opinion and practice differ in this respect. In the end, 
the setting should be checked by the taking of indicator 
diagrams, so that effects of incorrectible lost motion may be 
finally eliminated. 

With the dead-center points found the engine is placed 
on, say, head-end dead center, and the eccentric shifted until 
the valve is open to steam by the desired lead. The eccen- 
tric is then fastened in this position and the engine turned 
to the opposite dead center. Because of angularity of con- 
nections and of irregularities in valve and seat dimensions, 
it generally will be discovered that the valve is not now 
open to steam by the same amount as at the other end. 
If it is desired that it should be, the valve can be shifted 
on its stem about half of the distance by which it is out 
and the eccentric can then be swung about the shaft to take 
up the remaining distance. The effect should then be 
checked by putting the engine on the opposite dead center. 

Valves may be set for equal leads as above, or for equal 
cut-offs or for any sort of a compromise desired. In any 
case the procedure is about the same. The length of the 
valve stem is adjusted, then the eccentric position is 
adjusted, and then refinements are effected by small changes 
of both adjustments. Remember always, that changing 
the length of the valve stem changes events at opposite 
cylinder ends in opposite directions, while shifting the 
eccentric changes all events in the same direction. 

The effects of various adjustments are shown by the 



THE D-SLIDE VALVE 195 

indicator diagrams given in Fig. 128. These diagrams were 
taken from a small, slide-valve engine and serve very well 
to show the way in which the indicator discloses poor 
adjustments. 

PROBLEMS 

1. Given: angle of advance, 30°; throw of eccentric, If 
ins.; lead, r^ in.; maximum exhaust-port opening, 1| in.; find 
the steam lap, maximum opening to live steam, and the exhaust 
lap. 

2. Given: steam lap of | in.; lead of ^ in.; exhaust lap of 
f in.; and the angle of advance equal to 30°. Find the valve 
travel ( =2 X throw of eccentric) and maximum port opening to 
steam and to exhaust. 

3. An engine has an eccentric throw of If ins.; a steam lap 
of | in.; and a lead of ys hi- Compression begins at £ of the 
return stroke. Assume a connecting rod of infinite length and 
find the angle of advance, the exhaust lap, and the maximum 
port openings to steam and to exhaust. 

4. Given: valve travel, 3 ins.; steam lap, f in.; exhaust lap, 
\ in. ; and lead, f in. ; find maximum port opening, angle of advance, 
and piston positions at cut-off, release, compression, and admission 
for both ends of cylinder, with the length of the connecting rod 
equal to 4| times the length of the crank. 

5. It is required to build an engine having a steam-port opening 
of | in., a lead of ^ in., and a connecting rod four times the length 
of the crank. Cut-off must occur at § stroke and release at 95% 
of the stroke. Find the inside and outside lap, the throw of the 
eccentric and the fraction of stroke completed by the beginning 
of compression. 




CHAPTER XI 

CORLISS AND OTHER HIGH-EFFICIENCY ENGINES 

90. The Trip-cut-off Corliss Engine. The slide valve 
has certain limitations which can be partly, but never 
wholly, overcome. In most slide-valve gears, for instance, 
the various events occur more slowly than is desirable, 

and this is particularly 
true of cut-off. Ideal 
valves would open sud- 
denly to full opening 
when necessary and 
would close as suddenly 
at the proper time, and 
such action would give 
minimum throttling loss 
and rounding of corners 
of the diagram. Engines 
fitted with such ideal valves would therefore give indicator 
diagrams with maximum work area as shown by the 
dotted lines in Fig. 129, the full lines indicating the type 
of diagram obtained with the ordinary slide valve. 

Again, the simpler forms of slide valve involve the use 
of long ports connecting with the clearance space within 
the cylinder, thus adding greatly to the clearance surface 
exposed and to the cylinder condensation. These ports 
serve for both admission and exhaust, and their walls are 
therefore periodically cooled by the exhaust steam with the 
result that excessive condensation occurs during admission. 
Many attempts have been made to devise valve gears 
which should not be subject to the limitations of the 

196 



Fig. 129. 



HIGH-EFFICIENCY ENGINES 197 

simple slide valve. Some of these have resulted in the 
development of the more complicated slide valves de- 
scribed in the last chapter, but such designs generally leave 
much to be desired. One of the earliest and most success- 
ful solutions was made by Corliss, who developed what is 
known as the trip-cut-off Corliss gear. 

The long combined steam and exhaust ports are elimi- 
nated by the use of four valves, two for steam and two for 
exhaust. These are rocking valves and are located top and 
bottom, at the extreme ends of the cylinder, with their 
longitudinal axes perpendicular to those of the cylinder, 
as shown in Figs. 48, 49, and 50. The exhaust valves are 
located below so as to drain out water of condensation. 
Details of valves of this type are shown in Fig. 130. 

These valves may each be regarded as an elementary 
slide valve which has a cylindrical instead of a flat face, 
and which is oscillated about a center near the face instead 
of being reciprocated, i.e., oscillated about a center at an 
infinite distance. 

The valves are operated as shown in Fig. 131 by short 
links from a wrist-plate pivoted on the side of the cylinder 
and rocked back and forth about its center by means of an 
eccentric operating through the linkage indicated. The 
locations of the various pins and the lengths of the various 
links are so chosen that the valves travel at high velocity 
when opening and closing, that they open very wide, and 
that they close only far enough to prevent leakage and then 
remain practically stationary until about to open again. 
Throttling losses are thus decreased and wear caused by 
useless motion after closure is minimized. 

The opening of the admission valves in this gear is 
effected positively by the linkage already explained, but 
they are closed differently. For opening, the steam link 
rotates the bell crank B in Fig. 132 and thus raises the 
latch C. The hook on the end of one of the arms of this 
latch engages the steam arm which is fastened on the end 



X98 



STEAM POWER 



u 



I I 





HIGH-EFFICIENCY ENGINES 



199 




200 



STEAM POWER 



of a rod which is slotted into the end of the valve. The 
valve is thus drawn further open as the wrist plate revolves, 
until the tripping end D of the latch strikes the cam indicated 
by E. This throws the hook out of engagement and thus 
disconnects the valve from the driving mechanism. The 




Q o btea 

Fig. 132.— Details of Corliss Trip-Cut-off Gear. 

valve is closed by the action of a dash pot, one form of which 
is shown in Fig. 131. As the steam arm rises during the 
opening of the valve it draws up the plunger or piston of the 
dash pot, leaving a partial vacuum beneath it, and, when the 



20"x 48 Heavy Duty Corliss 
110 Lb. Steam 
K.F.M. 




Fig. 133. 

valve is released by unhooking of the latch, atmospheric 
pressure drives the plunger down and thus causes cut-off 
to occur. The action of a dash pot is found to be unsatis- 
factory when the speed of the engine exceeds about 125 
R.P.M. and most Corliss engines with trip-cut-off operate 



HIGH-EFFICIENCY ENGINES 201 

at still lower speeds. Under such circumstances the cut- 
off is very rapid as compared with the piston speed, and the 
diagram shows a comparatively sharp corner at this point. 
A set of diagrams obtained from a large Corliss engine 
operating at low rotative speed is given in Fig. 133, and it is 
obvious that little throttling occurs. 

Because of the low speed at which these engines operate 
the stroke can be made long with respect to the diameter 
without attaining a prohibitive piston speed. The economy 
mentioned in Chapter VII as resulting from the use of long 
strokes can thus be obtained in these engines. An idea 
of the saving in steam effected by the partial elimination 
of throttling and condensation losses by means of the 
Corliss gear can be obtained from the curves in Fig. 134 
(a) and (6), which give average performances. 

The position of the cam which determines the time 
at which cut-off occurs is controlled by the governor of 
the engine. When moved in the direction taken by the 
steam arm it causes cut-off to occur later. Variation of 
the point of cut-off is used in these and in most other engines 
to control the amount of work done per cycle in order that 
the engine may make available the quantity demanded at 
the shaft, as will be explained in a later chapter. It is there- 
fore desirable that the range of cut-off should be as great 
as possible, but it has been found very difficult to design 
trip-cut-off gears which will give a cut-off later than about 
0.4 stroke if steam and exhaust valves are operated from the 
same eccentric. Later cut-off causes poor timing of the 
exhaust events. 

This has led to the introduction of Corliss engines 
with two eccentrics and two wrist plates per cylinder. 
One set operates the steam valves and the other the exhaust 
valves. With this arrangement the range of cut-off is 
unlimited. 

91. Non-detaching Corliss Gears. Because of the low 
speed at which trip-cut-off Corliss engines are operated, 



202 



STEAM POWER 









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PROXIMATE STEAM CONSUMPTIO 

OF 

VARIOUS TYPES OF ENGINES 

(NON-CONDENSING) 

mple Single- Valve Throttling Engines (P = 100 lbs. 
mple Single- Valve Automatic Engines (P = 100 lbs. 
mple Four-Valve Automatic Engines (P = 100 lbs.) 
mdem and Cross-Compound Four- Valve and 

Corliss Engines (P = 100 lbs.) 
tndem and Cross-Compound Four- Valve and 

Corliss Engines (P = 125 lbs.) 
mdem and Cross-Compound Four- Valve and 

Corliss Engines (P= 150 lbs.) 
3ntz Engine P= 133 lbs.; 92.7° Superheat; 248.5 1 

R.P.M. 206. 
la-flow Engine, P= 150 lbs. 

na-flow Engine, P= 140 lbs.; Superheat, 110° F. 
fna-flow Engine, P= 150 lbs; Superheat, 300°. 


















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HIGH-EFFICIENCY ENGINES 



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APPROXIMATE STEAM CONSUMPTION 

OF 

VARIOUS TYPES OF ENGINES 

CCONDENSING) 

(1) = Simple Single-Valve Throttling Engines (P= 100 lbs.) 

(2) = Simple Single-Valve Automatic Engines (P = 100 lbs.) 

(3) = Simple Four- Valve Automatic Engines (P = 100 lbs.) 

(4) = Tandem and Cross-Compound Four- Valve and 

Corliss Engines (P= 100 lbs.) 

(5) = Tandem and Cross-Compound Four- Valve and 

Corliss Engines (P= 125 lbs.) 

(6) = Tandem and Cross-Compound Four- Valve and 

Corliss Engines (P= 150 lbs.) 
(») = Una-flow Engine (P= 140 lbs.). Superheat 100° F. 
(«') = Una-flow Engine (P= 150 lbs.), Superheat = 300° F. 















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204 



STEAM POWER 



they are necessarily large, heavy and costly and efforts 
have been made to design gears which shall possess the 
advantages of the original Corliss mechanism without 
the limitation as to speed. 

In many models the Corliss valves are retained and are 
located in the ends of the cylinder as just described or in 







Fig. 135. — Non-detaching Corliss Valves Located in Cylinder Head. 



the cylinder heads as shown in Fig. 135. In some the wrist 
plate and the connecting links are also retained, but in 
others they are eliminated. In all engines of this type the 
admission valves are closed positively, the closure being 
effected by the same linkage that opens the valves to admit 
steam. Quick action is obtained by the arrangement of 
the operating mechanisms, the centers of rotation and the 



HIGH-EFFICIENCY ENGINES 205 

lengths of links being so chosen that the valve travel is 
small when the valves are closed, that it is rapid when the 
valves are opening and closing, and that the valves remain 
practically wide open during most of the time that steam 
is being admitted. 

The advantages of small clearance and short and sepa- 
rate ports are attained in these arrangements and the 
operation of the valves is almost as perfect as that of the 
trip-cut-off gear. Engines fitted with these modified 
Corliss gears are operated at speeds considerably higher 
than those permissible with the older arrangement, and they 
may be classed with medium-speed engines. 

Engines of this type are generally known commercially 
as four- valve engines, but as this name applies equally well 
to the ordinary trip-cut-off gear and to others which will 
be described later, it is best to use some other designation. 
The term non-detaching Corliss engines seems to best 
describe them and is apparently gaining in favor. 

Non-detaching Corliss engines generally give diagrams 
intermediate between those obtained with the low-speed, 
trip-cut-off mechanism and those obtained from slide-valve 
engines with the simpler forms of valves, though the later 
designs very closely approximate the performances of the 
trip-cut-off Corliss engine. 

92. Poppet Valves. Attention has already been called 
to the fact that the use of highly superheated steam is 
very effective in lessening or even eliminating initial con- 
densation. Experience has shown that it is very difficult 
to make large valves with sliding surfaces, such as Corliss 
valves, work well with highly superheated steam. The 
large castings warp so that contact surfaces do not remain 
true and the lack of moisture which acts as a seal with 
saturated steam leads to excessive leakage. Difficulty 
has also been experienced with the lubrication of these 
sliding types of valves when using highly superheated 
steam. 



206 



STEAM POWER 




An old form of 
valve known as the 
poppet value has 
therefore been 
adopted by some 
builders as a solution 
of the difficulties met 
in the use of highly 
superheated steam. 
This form of valve in 
four-valve arrange- 
ment, combined with 
designs in which 
short ports and sym- 
metrical cylinder 
castings are used, 
yields very econom- 
ical engines which 
can be used safely 
with a degree of 
superheat prohibi- 
tively high in the 
case of the sliding 
and oscillating forms 
of valves. 




Fig, 136&. — Cross-section, Lentz 
Engine. 



HIGH-EFFICIENCY ENGINES 



207 



Sections of a modern type of poppet valve engine are 
shown in Figs. 136 (a) and 136 (6), and details of the 
admission valve and its operating mechanism are given in 
Fig. 137 (a) and (6). The valves are all double-seated 
(double-ported or double-beat), that is, they seat at both 
ends and are made hollow so that the steam passes both 
around the outside of the valve and through the valve 
as shown by the arrows in Fig. 137 (6). This results in 
large area for passage of steam and in quick opening and 




To Cylinder 



Fig. 137a. — Admission Valve and Operating 
Mechanism, Lentz Engine. 



Fig. 1376. 



closing, as in the case of gridiron valves, with small actual 
movement of the valve. 

The valves are opened positively by eccentrics opera- 
ting through cams and rollers as shown in Fig. 136 (6) and 
they are closed by springs as rapidly as the return motion 
of the cam permits. The eccentrics are mounted on a 
horizontal lay shaft which is located to one side of the 
engine, with its axis parallel to that of the latter, and which 
is driven by bevel gears from the crank shaft of the engine. 

Since this valve arrangement gives short steam and 
exhaust poits, permits the use of small clearance, and 



208 STEAM POWER 

gives fairly rapid opening and closing of valves with little 
throttling when open, it gives good economy when used 
with saturated steam. By adding superheat the economy 
is still further improved. The water rate of one of these 
engines is shown for one load in Fig. 134 (a). A simple, 
Lentz non-condensing engine is reported to have given 
a consumption of 16.13 lbs. of steam per horse-power hour 
with 92.7° superheat, and a pressure of 133 lbs., and this 
figure is materially lowered by compounding, higher super- 
heat, lower back pressure, etc. 

93. The Una-flow Engine. A very interesting modifica- 
tion of the steam engine, known as the Una-flow Engine, has 
been developed recently. The purpose of the design is to 
permit a great ratio of expansion in one cylinder while at the 
same time reducing the losses caused by initial condensation 
when such ratios of expansion are attempted in single cylinder 
engines of earlier types. 

Experience with the earlier types has shown that as cut- 
off is made earlier (ratio of expansion increased) in single 
cylinder engines, the steam economy is improved until best 
results are obtained with cut-off in the neighborhood of one- 
quarter stroke. With earlier cut-off the loss due to initial 
condensation overbalances the gain which should result from 
greater ratio of expansion so that a net loss ensues. 

The construction of the Una-flow Engine Cylinder is 
shown diagrammatically in Fig. 138. Steam enters the 
cylinder head at A and jackets the entire cylinder end as it 
flows toward the inlet valve located in the upper part of the 
head. The inlet valve is a double-seated poppet valve, the 
seats being indicated at B and C. Exhaust occurs through 
openings in the cylinder wall at the middle of its length, 
the engine piston serving as a valve to cover and uncover 
these openings. 

In operation, steam is admitted through the inlet valve 
as in other engines up to the point of cut-off. After cut-off 
this steam expands until the piston begins to uncover the 



HIGH-EFFICIENCY ENGINES 



209 



exhaust ports, thus effecting release. After completely 
uncovering these ports the piston returns and covers them 
again, the instant of complete closure corresponding to 




Fig. 138. — Section of Una-flow Engine Cylinder. 



beginning of compression in the ordinary case. Further 
motion of the piston compresses the trapped steam into the 
clearance space, and brings conditions back to those existing 
when admission starts. With an ideal arrangement the 
pressure in the clearance space at the end of compression 
would be just equal to that of the high-pressure steam out- 
side the inlet valve. 

It will be observed that the piston is so long that at either 
end of its stroke it fills all that volume of the cylinder between 
the clearance and the edges of the exhaust ports. The 
result is that the only part of the cylinder wall which is used 
in common for the cycles occurring at opposite ends of the 
cylinder is that part containing the exhaust openings or 
ports. It will also be observed that this part is strictly a 



210 STEAM POWER 

low temperature zone, coming in contact only with steam 
at exhaust temperature and pressure. 

This construction is almost equivalent to the placing of 
two single-acting cylinders end to end. It isolates each 
cylinder end in a sense, so that the thermal conditions 
resulting from the cycle carried through on one side of the 
piston are practically independent of those resulting from 
the cycle on the other side of the piston. In the ordinary 
form of double-acting engine the cylinder walls are common 
to both cycles for almost their entire extent with correspond- 
ing interrelation of thermal phenomena. 

Another important feature is the fact that high pressure, 
high temperature steam enters at the cylinder head while 
low pressure, low temperature steam resulting from expan- 
sion flows out of the cylinder at the point most distant from 
the point of entry. Steam flow is thus continuously in one 
direction and low-pressure steam does not sweep over parts 
which at the next admission will be bathed with high tem- 
perature steam. As a result of this construction, of the 
jacketing of the cylinder head with live steam on its way to 
the cylinder and of the long compression giving a compres- 
sion curve much nearer to the expansion curve than is com- 
monly obtained, the temperature distribution along the length 
of the cylinder is much more perfectly controlled than in the 
ordinary case. The ends of the cylinder tend to take and 
retain a temperature corresponding to that of the steam 
supply. The middle of length of the cylinder tends to take 
and retain a temperature corresponding to that of the 
exhaust steam. The length of wall between cylinder ends 
and center tends to assume temperatures grading from high 
temperature at the ends toward low temperature at the 
exhaust ports. 

As expansion of saturated steam occurs within a cylinder 
condensation occurs, work being done at the expense of the 
latent heat of vaporization thus released. This condensa- 
tion is distributed throughout the entire mass of steam and 



HIGH-EFFICIENCY ENGINES 



211 



at the metallic walls the water has a tendency to deposit. 
It should be noted that the temperature at which such water 
is formed must correspond at any instant to the temperature 
of the steam, and as the temperature of the steam decreases 
continuously as the pressure drops during expansion, the 
same thing must be true of the water formed during expan- 
sion. 

In the Una-flow design the cylinder head and walls near 
that head are kept hot by the steam jacket and if any water 
does deposit on them it is probably evaporated into steam 



R 


Steam 


pressure 


at the Throttle 150 lbs. Vacuum 27" 




p 




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W 

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60 



80 90 100 110 120 
Indicated Horse Power 



130 



140 150 



Fig. 139.— Results of Tests Made with a " Una-flow Engine. 



again. The engine is assumed to operate in this way and it 
appears to be true that the greater part of the water resulting 
from expansion concentrates near the exhaust ports and is 
swept out during the release of steam. Such action is 
indicated diagrammatically in Fig. 138 by the gradation of 
the stippling within the cylinder. 

At first sight it might appear that heat transferred 
from jacket to steam within the cylinder would surely result 
in loss but it should be noted that a large part of the steam 
within the cylinder which is in contact with steam-heated 
surfaces probably remains in the cylinder and constitutes 
the clearance steam so that heat transferred to it during 



212 



STEAM POWER 





expansion does not pass out 
through the exhaust valve to 
such an extent as it would in 
anengineoftfaeordinarytype. 

The long compression 
stroke, with increasing tem- 
perature and pressure, also 
has a tendency to evaporate 
any moisture remaining on 
the walls. This is certainly 
true for that part of the 
metal near the cylinder head. 
The entering steam thus 
comes in contact with hot, 
dry walls and with hot, dry 
steam and initial condensa- 
tion is greatly reduced, if 
not eliminated. 

Remarkably low steam 
consumptions have been ob- 
tained with Una-flow en- 
gines, even when simple en- 
gines have been used to 
expand steam from compar- 
atively high pressure to a 
high vacuum. Results ob- 
tained with one engine under 
different conditions are 
shown in Fig. 139. 

The mechanical con- 
struction of an 18 X 20-inch 
single cylinder Una -flow en- 
gine is shown in Fig. 140, a 
section through the exhaust 
ports being shown to the 
right of the figure. 



HIGH-EFFICIENCY ENGINES 213 

94. The Locomobile Type. In the effort to improve 
the ecomony of small steam plants the Germans developed 
a form of plant now known as the Locomobile Type. The 
name came from the fact that these plants, as originally 
made, were mounted on wheels and intended for portable 
use by agriculturists and contractors. Their economy in 
the use of fuel proved so great that they have since been 
built for stationary use in sizes running well toward 1000 
horse-power per unit. 

A locomobile of American construction known as the 
Buckeye-mobile is illustrated in Fig. 141, which shows a 
longitudinal section of the plant. The tandem compound 
engine is mounted on top of an internally fired boiler with 
the engine cylinders located in the flues which lead the 
products of combustion away from the boiler. 

The steam generated in the boiler is passed through 
a superheater suspended in the smoke box. The flow of 
steam is from the rear toward the front of this superheater 
(counter flow) so that the hottest steam comes in contact 
with the hottest gas. The steam then passes through a 
pipe contained within the flue to the high-pressure cylinder, 
which is jacketed by the hot flue gases and in which the 
loss of heat to metal is thus minimized. From the high- 
pressure cylinder the steam passes to a receiver contained 
in the smoke box, the receiver serving as a reheater to 
evaporate any condensate exhausted from the first cylinder 
and to superheat the steam admitted to the low-pressure 
cylinder. From the low-pressure cylinder, the steam 
passes through a feed-water heater in which it raises the tem- 
perature of the boiler feed and then it passes to atmosphere 
or to a condenser. Boiler-feed pump and condenser pump, 
if used, are also integral parts of the plant, being driven 
directly from the main engine. 

It' will be observed that every precaution is taken to 
guard against initial condensation, and to minimize loss 
of heat in flue gases and in exhaust steam leaving the 



214 



STEAM POWER 




HIGH-EFFICIENCY ENGINES 215 

plant. The high economies achieved are due to such facts 
alone. 

Small plants of this type have given an indicated horse- 
power hour on a little over one pound of coal when operated 
condensing, whereas the best large compound reciprocating 
engine plants seldom do better than about 1.75 lbs. of coal 
per I.h.p. and often use 2 or more pounds when operated 
condensing. 



CHAPTER XII 
REGULATION 

95. Kinds of Regulation. There are two distinctively 
different kinds of regulation referred to in connection 
with reciprocating steam engines, one of which may be 
called fly-wheel-regulation and the other governor-regula- 
tion or governing. 

The regulating effect of the fly-wheel has already been 
referred to. The turning effort exerted at the crank pin 
by the action of steam on the piston or pistons of an engine 
is not constant, and the angular velocity of the engine shaft 
is therefore constantly varying during each revolution. 
It is the function of the fly-wheel to damp these variations 
so that they do not exceed the allowable maximum for 
any given set of operating conditions. The efficiency of 
the fly-wheel in this respect is measured by the coefficient 
of fly-wheel regulation 8 W which is defined by the equation 

-. 'max ' mln /aa\ 

0w = y , \pO) 

in which 

Fmax = maximum velocity attained by a point on 

fly-wheel rim or other revolving part; 
Fmin = minimum velocity of the same point, and 
F = mean velocity of the same point 

= - approximately. 

Governor-regulation is absolutely different. Its function 
is to proportion the power made available to the instan- 
taneous demand. The fly-wheel takes care of variations 

216 



REGULATION 



217 



occurring during the progress of one cycle, while governor 
regulation varies the work value of successive cycles. 

96. Governor Regulation. If the effect of engine 
friction be neglected, the power delivered at the shaft of 
the engine will vary directly with the indicated horse- 
power. Such an assumption is accurate enough for the 
discussion which follows. 

The indicated horse-power of a given engine is deter- 
mined entirely by the value of the mean effective pressure 
and the number of cycles produced in a given time, since 
these are the only variables in the formula for indicated 
horse-power. The power made available by an engine 
might therefore be varied by varying the mean effective 





Fig. 142.— Throttling 
Governing. 



Fig. 143.— Cut-off 
Governing. 



pressure, or by varying the number of cycles produced in 
a given time, or by a combination of both processes. 

All of these possibilities are used. In ordinary station- 
ary power plants the mean effective pressure is generally 
varied. In the case of pumping engines, working against 
a constant head, but required to deliver different quantities 
of water at different times, the number of cycles per minute 
is generally altered by changing the speed at which the 
engine operates. In locomotive and hoisting practice 
both the number of cycles per minute (speed) and the 
mean effective pressure are varied as required to meet 
the instantaneous demands. 

These variations may be effected manually as by the 
driver of a locomotive, in which case the engine may be 
said to be manually governed. Or, they may be brought 



218 STEAM POWER 

about mechanically, as in the case of most stationary power- 
plant engines, in which case the engine may be said to be 
mechanically governed. In some instances a combination 
of manual and mechanical governing is used. 

97. Methods of Varying Mean Effective Pressure. The 
mean effective pressure increases and decreases with the 
area of an indicator diagram of constant length, so that 
the mean effective pressure can be changed by any method 
which will change the area of the diagram. Two methods 
are in use and they are illustrated in Figs. 142 and 143. 
The first causes a variation in area by changing the value 
of the initial pressure. This is generally done by chang- 
ing the opening of a valve in the steam line just outside of 
the steam chest. It is called throttling governing, and the 
valve is called a throttling or throttle valve. The latter 
name is also commonly used for the valve located near the 
engine, which is used to shut off the supply of steam entirely 
when the engine is not in operation. 

The second method, illustrated in Fig. 143, is known 
as cut-off governing. The variation of cut-off determines 
the amount of steam admitted to the cylinder per cycle 
and is used to measure out the quantity required for the load 
which happens to exist at any instant. Cut-off governing 
is used on most modern stationary engines and is exclusively 
used in large reciprocating engine power plants. 

98. Constant Speed Governing. Most engines used for 
such purposes as the operation of mills and the driving of 
electrical and centrifugal machinery are required to run 
at practically constant speed irrespective of the load. They 
are furnished with mechanical governors which so regulate 
the power made available that there shall never be any 
appreciable excess or deficiency which would respectively 
cause an increase or a decrease in speed. 

These mechanical devices always contain some sort 
of tachometer which moves whenever the speed of the engine 
exceeds or falls below the proper value. The tachometer 



REGULATION 219 

is so connected to the valve gear that it decreases the 
power-making ability of the engine whenever the speed 
starts to increase and it increases the power-making ability 
if the speed drops. 

Since the valve gear must have a different position 
for each load in order that it may throttle or cut off as 
necessary to suit that load, it follows that the tachometer 
which controls the position of the valve gear must also 
have different positions for different loads. But tachom- 
eters assume positions dependent on speed, and therefore 
different loads can only be obtained if the tachometer and 
the engine to which it is connected operate at different 
speeds for different loads. 

Constant-speed governing is therefore an anomaly. 
The device which is supposed to maintain constant speed 
irrespective of load must be operated at different speeds, 
as the load varies, in order that it may maintain the valve 
gear in the different positions required to handle the differ- 
ent loads. All so-called constant-speed engines have their 
highest speed when carrying no load, and the speed gradually 
decreases to a minimum as the load increases to a maxi- 
mum. The total variation is generally between 2 and 4%. 

The efficiency of a governor in this respect is measured 
by means of the coefficient of governor regulation, 8 G , 
which is defined by the equation 

n 2 — ni , 

8g= , (67) 

n 

in which 

ri2 = highest rotative speed attained by the engine; 

n\ = lowest rotative speed attained by the engine, and 

n = mean speed 

n 2 +ni . . 

= — — approximately. 

99. Governors. The mechanical devices which are 
used for controlling the power-making ability of an engine 



220 



STEAM POWER 



as described above are known as governors. There are 
many varieties and only a few of the more prominent can 
be described. 

(a) The Pendulum Governor. One of the earliest forms 
of governor used on steam engines is illustrated in Fig. 
144. It is often called a 
fly-ball governor. This 
governor is driven by 
gearing, chain or belt 
from the engine, and the 
weights assume some 
definite position for 
each different speed, thus 
drawing the collar to 
different positions. The 
valve gear is connected 
to this collar and is 
moved correspondingly. 

A similar governor 
is shown in Fig. 131, 
which also indicates the 

way in which the collar is connected to the valve gear in 
the Corliss type of engine. The governor rods are moved 
as the collar moves and they in turn alter the position 
of the knock-off cam, and thus vary the time at which 
cut-off occurs. As the speed increases due to a decrease 
of load, the governor weights and collar move up, and this 
shifts the cams so as to produce earlier cut-off and decrease 
power-making ability. 

(6) Shaft Governors. On medium- and high-speed 
engines fitted with some form of slide valve it is found best 
to use what are known as shaft governors. They are gen- 
erally carried within the fly-wheel of the engine, operate in 
a plane passing through the rim of the wheel at right angles 
to the shaft, and operate upon the eccentric in such a way 
as to vary the cut-off with speed (and load) changes. 




Fig. 144. 



REGULATION 



221 




Fig. 145. 



One simple form of such a governor is shown in Fig. 
145. The eccentric is not mounted directly upon the 

engine shaft, but is carried 
by a pin P in the fly-wheel 
and is slotted so that it can 
swing back and forth across 
the shaft, about P as a center. 
Its position at any time is 
determined by the position 
of the governor weights W, 
which draw the eccentric 
down (in the figure) as they 
move out. 

The center of the eccentric 
is indicated by a heavy dot 
in the figure, and it will be seen that this center would 
travel in the arc of a circle about P, as the weights moved. 
If the path of the eccentric center is drawn on a Bilgram 
diagram, it will be found that this motion is equivalent to 
decreasing the length of the eccentric crank and increasing 
the angle of advance, resulting in earlier cut-off as the 
weights move out with increasing speed and decreasing 
load. Other events will also be changed as the eccentric 
swings, and some of these changes are occasionally unde- 
sirable. 

Numerous designs have been developed in which the 
eccentric is so guided as to produce various sorts of rela- 
tions between the different steam and exhaust events. 
All can be divided into two classes, those in which the 
eccentric swings about a fixed center variously located, and 
those in which the center of the eccentric is guided to 
move in a straight line. All can be studied by plotting 
the path of the eccentric center (path of Q) on the Bilgram 
diagram. 

The Rites Inertia Governor is a form of shaft governor 
so designed as to act very quickly with change of speed, 



222 



STEAM POWER 



and to be very powerful, so that it can shift heavy parts. 
It is shown in place in the wheel in Fig. 146. With changes in 
speed it acts like a governor of the type just described, 
swinging (with increasing speed) about a fixed point P in 
the wheel as its center of gravity G moves outward under 
the action of the centrifugal effect C and against the 
action of the spring. This motion shifts the center of the 




Fig. 146. 



eccentric from E toward e, giving the desired variation in 
cut-off. 

Superposed upon this action is that of inertia. Assume 
the wheel and governor to be rotating clockwise at a given 
constant speed. If the engine speed is suddenly increased, 
the wheel will move faster, but the governor bar will tend 
to continue rotating at the same speed because of its inertia. 
It will thus lag behind the wheel, rotating about P and bring- 
ing about an earlier cut-off. The position thus assumed 
will later be maintained by centrifugal effect if the new speed 



REGULATION 223 

is maintained. The particular advantage resulting from 
using inertia in this way is speed of action. In many forms 
of governor the inertia of the moving parts actually resists 
the efforts of the governor to assume the new position 
required by changed load and speed, whereas in this form 
the inertia of the governor parts is used to increase the speed 
with which they move to the new position. 



CHAPTER XIII 

THE STEAM TURBINE 

100. The Impulse Turbine. One of the oldest of modern 
water wheels is the tangential or impulse wheel shown 
diagrammatically in Fig. 147. Water flowing from a 
reservoir above the wheel passes through a nozzle and the 




Fig. 147. — Tangential or Impulse Wheel. 

jet, moving at high velocity, strikes buckets on the rim 
of the wheel and causes the latter to revolve. Theoretically 
the velocity of the water in the jet would be 

v = V2gh feet per second, .... (68) 
in which 

g = gravitational constant, 32.2, and 
h = head in feet as shown in the figure. 

The kinetic energy possessed by the moving water would 
be 



K = 



w v* 

7"2' 
224 



(69) 



THE STEAM TURBINE 



225 



in which iv represents pounds of water discharged per second 
and g and v have the same meanings as above. 

If the buckets of the wheel could reduce the velocity 
of the water to zero they would absorb all of this kinetic 
energy and (assuming no losses within the buckets and the 
bearings of the wheel) would make all of it available at 
the shaft for the doing of useful work. 

Any fluid moving at velocity v and striking buckets in 
the form of a jet would possess kinetic energy in quantity 
given by Eq. (69) and would drive the wheel in the same 
way. Steam might therefore be used instead of water 
with exactly the same results, and steam is so used in what 
are known as impulse steam turbines. 

Experience shows that steam will flow at high velocity 

from any opening made in the 
s^> steam space of a boiler or 
from any open-ended pipe con- 
nected to such a boiler. This 
is commonly said to be due 

— ^ _j | to the high pressure within 

C^M^,,^^ gJL, the boiler, the spectator pic- 

turing the process as the 
driving out of part of the 
steam by the high-pressure steam within the boiler, just as 
though the part leaving were a solid piston and were driven 



vS^^^Pf 



^ 



Fig. 148. 



Fig. 149. 



out as is the piston of an engine during 
admission, as shown in Fig. 148. 

An hydraulic analogy is given in Fig. 
149. The vessel shown is supposed to be 
fitted with a piston, and it is assumed to be possible to exert 
any desired pressure upon the piston. Any such pressure 
exerted is the exact equivalent of some given head of water 
and the resultant jet velocity would be given by Eq. (68) 
by substituting for h the head in feet equivalent to the 
pressure exerted upon the piston. 

When an " elastic " fluid such as steam is being con- 



226 STEAM POWER 

sidered it is, however, necessary to take account of other 
factors. The steam within the boiler exists at a high 
pressure ; after issuing it exists in the atmosphere at a lower 
pressure. But low-pressure steam contains less heat than 
does steam at high pressure, and this difference must exist 
in some form, as it is energy and could not possibly have been 
destroyed during the flow. 

Experiment shows that steam after flowing into the 
atmosphere from a boiler in this way has exactly the same 
characteristics as though it had expanded adiabatically 
behind a piston through the same temperature range, ex- 
cepting for the fact that it has a very high velocity, which 
it would not possess if expanded behind a piston. Experi- 
ment further shows that, if small losses be neglected, the 
kinetic energy possessed by a jet of steam is exactly equal 
to the energy which would be turned into work if that steam 
acted on a piston as in an ordinary engine. 

A complete picture of the process of flow can then 
be made by assuming the steam flowing out in the form of 
a piston driven by high-pressure steam, as before, and adding 
to this the idea that this piston expands adiabatically as 
it travels from the region of high to that of low pressure. 
This expansion liberates heat contained within the piston 
or plug of steam and this heat is used in imparting addi- 
tional velocity to the moving steam which is giving up this 
heat. 

The result of using such a jet upon a theoretically 
perfect tangential or impulse wheel would be to rob the 
jet of all this energy. But the energy possessed per pound 
of steam in the jet is just the same as that shown under the 
upper lines of a complete expansion cycle using one pound 
of steam. The area under the upper horizontal line of the 
PF-diagram of the cycle as shown in Fig. 21 may be assumed 
to represent the work done upon one pound of steam (flow- 
ing out) by another pound which is being evaporated and 
pushing out the first in order to make room for itself. The 



THE STEAM TURBINE 



227 



area under the expansion curve in the PF-diagram repre- 
sents the energy converted into velocity energy by the adia- 
batic expansion of the flowing steam. The lower hori- 
zontal line represents the negative work during condensa- 
tion to water at the lowest pressure and temperature, and 
the left-hand line represents the pumping of this water back 
into the boiler and the raising of its temperature to the value 




tDiarihragm 



Fig. 150. — Early Form of Impulse Turbine. 



maintained within the boiler. The complete expansion cycle 
is therefore the cycle upon which the impulse steam turbine 
operates and, as a matter of fact, it is the theoretical cycle 
of all steam turbines. 

The ideal impulse turbine would therefore be acted 
upon by a jet which possessed available kinetic energy 
represented by the area of the complete expansion cycle. 
If the buckets could entirely remove this energy, that is, 
could reduce the velocity of the jet to zero, the same amount 



228 



STEAM POWEK 



of energy could theoretically be made available at the shaft 
of the turbine. 

An example of a simple form of impulse steam turbine 
is given in Fig. 150, in which the essential parts of an early 
form of Kerr turbine are shown. The wheel, the diaphragm 
and nozzles are all inclosed within a casing. The space 
on one side of the diaphragm is connected to the steam pipe 
and that on the other is in communication with the space 
into which the exhaust steam is to be exhausted. 

Another form of impulse turbine is shown in Fig. 158. 
It will be described later. 

101. Theoretical Cycle of Steam Turbine. It was shown 
in the preceding section that the steam turbine operates 
on the complete expansion cycle. If a turbine could remove 
from the steam passing through it and convert into mechan- 
ical form all the energy which it is theoretically possible to 
convert it would therefore make available mechanical energy 
represented by the area of the PF-diagram of the complete 
expansion cycle. The area of the corresponding T<p- 
diagram would show the 
same quantity measured in 
thermal units. The theory 
of the steam turbine can 
therefore be studied by 
means of these two dia- 
grams. 

In Fig. 151 is shown the 
T^-diagram of the complete 
expansion cycle for several 
different conditions. The 
figure abed represents con- 
ditions when the steam is 

dry and saturated at the beginning of the adiabatic ex- 
pansion cd. Constant quality lines are designated by x 
and %' . It is obvious that by the time the steam has 
expanded down to the pressure at d it will have a quality 




Fig. 151. 



THE STEAM TURBINE 229 

less than unity. If, therefore, it be in the form of a jet 
issuing from a nozzle and having a high velocity by virtue 
of its adiabatic expansion, the jet will really be a mixture 
of steam and water. 

If the steam be superheated at constant pressure as shown 
by ce before passing through the nozzle, it is evident from the 
figure that the jet issuing from the nozzle will contain 
less water than in the preceding case, because the condition 
of the material in the jet after adiabatic expansion will be 
as shown at / instead of as shown at d. The cycle in such 
a case would also be larger by an amount indicated by the 
area cefd, representing just that much more heat converted 
into mechanical energy per pound of steam or other unit 
for which the diagram happened to be drawn. 

If superheating had been carried to the point indicated 
by g before expansion, the jet would obviously issue from 
a nozzle in the form of superheated steam as shown by the 
point h in the figure. In that case the cycle would be 
abcgha, and superheat would have to be removed from the 
low-pressure steam to bring it to the conditions indicated 
at i before condensation could begin. 

If desired, the PF-diagrams for such cycles can be drawn 
very easily. The line 6c, or be or bg is a horizontal line in 
the PT-diagram. The line ha is similarly horizontal and the 
line ab is vertical. The adiabatic expansion is represented 
by a curved line in the PF-diagrams, but can be drawn 
easily because the necessary data are obtainable from the 
T<£-diagram, in which this expansion is represented by a 
straight line. 

ILLUSTRATIVE PROBLEM 

Draw the PF-diagram for a steam turbine receiving one pound 
of steam at a pressure of 200 lbs. absolute, with a tempera- 
ture of 500° F. and exhausting against a pressure of 0.5 lbs. 
absolute. 

First, locate on a T^-chart for steam the point representing 
the condition of steam at 200 lbs. pressure with a temperature 



230 



STEAM POWER 



of 500° F., and draw a vertical line extending downward 
until it cuts the horizontal temperature line corresponding to 





































^ 












































































































• 
/ 




































/ 


















/ 
















/ 


















/ 




J 


3 




__'••> 




»-"' - 






a 



3 



O pE< 



0.5 lb. pressure. This is practically at 540° F. absolute, or about 
80° F. 



THE STEAM TURBINE 231 

Second, take from the steam table the volumes of one pound 
of steam at, say, 200 lbs., 140 lbs., and 100 lbs. absolute pressure 
when superheated to the values shown by this vertical line. These 
will be about 2.75 cu.ft., 3.58 cu.ft., and 4.67 cu.ft., respectively. 
Plot these volumes with corresponding pressures on a PF-chart 
as shown in Fig. 152. 

Third, take from the TV-chart the pressures at which the 
vertical line intersects different volume lines in the wet steam 
region and plot volumes against pressures on the PF-chart. 

Fourth, draw a smooth curve, as shown, through all points 
so determined. 

Fifth, draw horizontal top and bottom lines and a vertical 
line at the left of the diagram. This vertical line should be to 
the right of the pressure axis by an amount representing the 
volume of one pound of water, but the volume is so small that 
it cannot be plotted to any ordinary scale. 

102. Nozzle Design. It was stated In preceding sec- 
tions that the energy which would be converted into work 
by the introduction and adiabatic expansion of steam 
behind a piston is converted into kinetic energy when steam 
flows out of an orifice or nozzle and that an ideal impulse 
turbine could absorb all this kinetic energy from the jet, 
bringing it to rest and making the energy available in the, 
form of useful power at its shaft. It is, therefore, of interest 
to determine the velocity which a jet will acquire under 
different conditions. 

This could be done by evaluating the area of a diagram, 
such as that of Fig. 152, and then putting this value m 
place of K in Eq. (69) and solving for v, but it can be done 
much more accurately and expeditiously in other ways. 
The heat energy which can be converted into kinetic 
energy of the moving jet and which can later be con- 
verted into useful work by the turbine wheel is represented 
by the area enclosed within the lines of the complete 
expansion cycle when drawn on the T^-diagrams. That 
is the area abed in Fig. 153, for instance, for the case of 
wet steam at the beginning of expansion. But this area 
is equal to that representing the heat supplied minus 



232 



STEAM POWER 



that representing the heat rejected, that is, Q1-Q2, so 
that 

K(inB.t.u.)=Qi-Q2. .... (70) 

The values of Qi and Q 2 can be found very readily by 
plotting the points c and d upon a T</>-chart for steam and 
observing the constant heat lines upon which they fall, 



a 


b 

1 ' 


c 


d \ 


/ 


i 












\ > 






■2 










1 \ 


' 



Fig. 153. 



or they can be obtained even more conveniently from what 
is known as a Mollier Chart for steam. In this chart, 
entropy above 32° F. is plotted against heat above 32° F. 
as shown in Fig. 154. An adiabatic expansion on this chart 
is shown by a horizontal line, since this shows a constant 
entropy change just as a vertical line on the Tcf> chart shows 
a constant entropy change. 

If a point is found in this chart giving conditions corre- 
sponding to those at point c in Fig. 153, the value of Q x 



THE STEAM TURBINE 



233 




«2 Mo.T}ua »> 



234 STEAM POWER 

can be read directly under that point on the horizontal 
axis. A horizontal line drawn from that point to the 
terminal-pressure line will give the point corresponding to 
d of Fig. 154 and the value of Q2 can be read on the hori- 
zontal axis immediately below that point. The difference 
between the two readings gives the value of the kinetic 
energy K or of the mechanical energy which an ideal tur- 
bine could make available, but the expression will be in 
British thermal units and not in foot-pounds. 

This value of the kinetic energy, i.e., K = Q\ — Q2, may 
then be placed in Eq. (69), giving, 

X = 778(Qi-Q 2 )=£ft.-lbs., . . . (71) 
since Qi and Q2 refer to one pound of steam, or 

K = 77MQ-Q 2 )=7rft.-lbs., • • • (72) 

when w represents the number of pounds of steam flowing 
per second. 

Solving cither Eq. (71) or Eq. (72) for v gives, 

v = V77SX2g(Q 1 -Q 2 ) 

= V / 778X64.4(Q 1 -Q 2 ) 

= 22WQi-Q 2 feet per second. . . (73) 

The design of a nozzle consists simply in choosing such 
sections that the desired amount of steam may flow through 
it with the desired pressure drop, as the velocity obviously 
is determined by that pressure drop. This is very con- 
veniently done by working in terms of one pound of steam, 
since all formulas and charts are generally given on that 
basis, and then multiplying the cross-sectional areas found 
by the number of pounds of steam required. 

Assume for instance that it is desired to design a nozzle 



THE STEAM TURBINE 235 

to pass one pound of steam per second with an initial pres- 
sure of 100 lbs. per square inch abs., and a terminal 
pressure of 60 lbs., the steam being initially dry and satu- 
rated. 

The Mollier chart shows that Qi is equal to about 1187 
B.t.u. per pound of steam, while Q2 is equal to about 1147 
B.t.u. The velocity with which a jet would issue from a 
theoretically perfect nozzle under these conditions may 
then be found by using Eq. (73). This gives 

v = 22Wl 187 -1147 
= 1416 feet per second. 

The shape of the entrance end of the nozzle is generally 
made such that the steam will enter it without great dis- 
turbance and the shape beyond that point is determined 
by methods which will be explained below. The cross- 
section of the discharge end must be such as to pass the 
required quantity at the velocity found above to be equal 
to 1416 feet per second. This is easily done by deter- 
mining the volume of steam discharged. 

Drawing the adiabatic expansion on the T<£-chart will 
give the quality at the end of the expansion; or, the quality 
can be determined by finding what quality a pound of 
steam at 60 lbs. pressure must have to give it a heat content 
of 1147 as found above. With the quality known the ter- 
minal volume per pound can be found by multiplying the 
quality by the specific volume at terminal conditions. 
Thus for the case under discussion the quality will be 
about 96.7% and as the specific volume at 60 lbs. is 7.17 
cubic feet, the volume to be passed per second, per pound 
of steam is 0.967X7.17 = 6.94 cu. ft approximately. If the 
velocity is 1416 feet per second the area per pound of steam 
must be 6.94-^-1416 = 0.0049 sq.ft. 

The exact shape of the nozzle is determined by deciding 
upon the way in which pressure, or velocity, or volume 



236 



STEAM POWER 



shall change as the steam passes through' it. Suppose, for 
instance, that a nozzle is to be constructed of the length 
shown by ab in Fig. 155, and that the pressure is to vary 
along its length as shown. Assume also that the nozzle 
is to pass 10 lbs. of steam per second. Taking initial 
pressure as 100 lbs. and terminal as' 60 lbs., the conditions 



a 
^100 



90 



80 



70 



2 60 

























*^«^ 






























f> 































s 

/ 


















i, 


7 








?2 


























\t> 


b 


















x3 


* / 


















/N£ 


^ 


tfj, 










/ 








<^v? 










/ 


















fVa 


riati 


3ii o: 


Vel< 


>city 



























1400 
1200 
1000 
800 
600 
400 
200 



Length of Nozzle 
Fig. 155. — Nozzle Design. 

will be the same as in the problem above. The discharge 
area will have to be 10X0.0049 sq.ft. or 0.049 sq.ft. 

The area at the plane X2 must be that required to pass 
the steam when it has the velocity resulting from expansion 
from 100 down to 64 lbs., just as though the nozzle ended at 
that point. This can be found just as the terminal area 
was found above. Similarly the sections at x\ and x can 
be found by figuring velocity and area for expansions to 



THE STEAM TURBINE 



237 



74 and 90 lbs., respectively. If the various areas required 
are determined in this way, the nozzle will have a longi- 
tudinal section about as shown by the dotted lines in the 
figure and the variation of velocity will be about as shown 
by the curve. 

If the shape of a nozzle is determined in the same way 
for a case in which the terminal pressure is less than about 
0.58 of the initial pressure, the nozzle will be found to have 

a very different shape. This is 
shown in Fig. 156. The nozzle 
is known as an expanding nozzle 
and the smallest section is known 
as the neck. The pressure P» 
in the neck is always equal to 
about 0.58 Pi and the velocity 
in the neck is always equal to 
just over 1400 feet per second. 
It is therefore the section at 
the neck which determines the 
quantity of steam which a nozzle 
will discharge if expanding to a 
pressure equal to or lower than 
0.58 Pi. 

103. Action of Steam on 
Impulse Blades. It has been 
stated that the steam acting in an impulse type of turbine 
delivers energy to the wheel of the turbine by giving up 
its kinetic energy. In an ideal turbine the steam jet would 
be brought to rest and would thus give up all of its kinetic 
energy. 

In real turbines it is impossible to bring the jet to rest, 
as practical design problems prevent it. There is there- 
fore always a loss in real machines because of the residual 
or terminal velocity of the steam as it leaves the wheel. 
Thus let the black section in Fig. 157 represent the section 
of a bucket or blade sticking out radially from the rim of a 




Fig. 156. — Expanding Nozzle. 



238 



STEAM POWER 



wheel, the wheel revolving about the axis indicated by the 
dot dash line but located behind the plane of the paper, 
see Fig. 158. If minimum loss by eddying is to be experi- 
enced at the point at which the 
steam jet enters the blade, the 
jet must enter the blade along a 
tangent to the curve of the in- 
side at the entrance edge. This 
direction is shown by the line 
marked v r in the figure. 

Were the bucket stationary, 
the steam jet would move as 
shown hyv r , but as the bucket 
moves ahead, and, so to speak, 
runs away from the jet, the 
steam must really travel in a 
direction such as that indicated 
by v a in order to strike the 
bucket in the direction indicated 
by v r . The conditions governing 
the flow of steam into a bucket 
are the same as those governing the speed with which and 
direction in which an individual runs toward and jumps 
upon a moving vehicle. He will experience least shock 
when he is moving ahead at the same rate as the vehicle 
at the instant when he gets on board. His motion must 
therefore be made up of two, one toward the vehicle and 
the other in the direction of the vehicle's travel. 

In the case of steam flowing onto a blade as shown in 
Fig. 157, the various velocities are so related that when 
drawn to scale the real or absolute velocity of the steam, 
v a , and the real or absolute velocity of the blade, v b , form 
two sides of a triangle of which the closing side represents 
v r , the velocity of the steam relative to the bucket. The 
value and direction of v T is obviously obtained from v a by 
geometrically subtracting the velocity of the bucket. 




Fig. 157. 



THE STEAM TURBINE 239 

After entrance, the steam flows around the inner curve 
of the blade and is finally discharged with the same rela- 
tive velocity as that with which it entered, and at an angle 
set by the tangent to the inner curvature of the discharge 
edge of the blade as shown by v R . But, since the steam 
has been moving ahead with the same velocity as the 
bucket during the entire time that it was in contact with 
the bucket, it is also moving ahead with a velocity Vb when 
it leaves the wheel. Its real or absolute velocity is then 
v A , which is found by combining v R and v b as shown in the 
figure. 

The kinetic energy possessed by the jet when entering 

wv 
the blade is equal to ~~ ft.-lbs., and that which it possesses 

WVa 

when leaving is — = — ■. Obviously, the energy removed 

VOV it VOV a 

while passing over the blade is -~ — •. If the blade 

were theoretically perfect, it would be so constructed that 
v A 2 would be zero and all of the kinetic energy would then 
be removed. This is practically impossible in a real mechan- 
ism, and there is always a loss due to the residual velocity 
v A . The best that can be done is to so choose the angle 
of jet and blade, and the velocity of blade with respect 
to the steam, that the actual numerical value of v A is made 
as small as possible. 

Designs usually work out in such a way that this occurs 
when the blade velocity is equal to about 0.47 of the abso- 
lute velocity of the steam jet. 

104. De Laval Impulse Turbine. The expanding nozzle 
already described was first used by De Laval in an impulse 
type of turbine. The essential elements of this device are 
shown in Fig. 158. The nozzles are arranged at such an 
angle to the plane of the wheel that the steam jets strike 
radially arranged blades at the proper angle to enter without 
much loss. The blades deflect the jets as shown and 



240 



STEAM POWER 



absorb the greater part of their kinetic energy, so that 
the steam flows away from the wheel with low absolute 
velocity. 

As many nozzles are used as are required to make avail- 



steam In 



Nozzle 




Turbine Shaft 



. tNozzles- 
Fig. 158. — Single Stage, De Laval Impulse Turbine. 

able the amount of energy desired at full load, and pro- 
vision is made for shutting off one or more nozzles by hand 
when conditions do not warrant the use of all. Governing 
for ordinary variations of load is effected by throttling 
the steam flowing to the nozzles in use, thus altering the 
initial pressure as necessary. 



THE STEAM TUEBINE 241 

A section through the wheel and casing of such a tur- 
bine " direct connected " to a centrifugal pump is given 
in Fig. 159. The steam flows into the live steam space 
through a throttle valve controlled by the governor; the 
valve and connections are not shown in the illustration. 
From the live steam space the steam flows through nozzles 
not shown, and into the exhaust steam space, thus acquir- 
ing a high velocity. The buckets of the wheel are located 
just in front of the discharge ends of the nozzles and the 
steam moving at high velocity must pass through them 
before moving on toward the exhaust outlet. 

105. Gearing and Staging. It has been stated that the 
most efficient operation with ordinary designs is obtained 
when the blade speed is equal to about 0.47 of the absolute 
steam velocity or, roughly, half the velocity of the imping- 
ing jet. To get high economy in the use of steam, large 
pressure drops are used and very high jet velocities result. 
When the buckets of a turbine are operated at peripheral 
speeds equal to half these jet velocities one of two diffi- 
culties is often met. The stresses induced in the wheel 
structure by centrifugal effects become so high as to offer 
serious difficulties in design, or the rotative speed of the 
unit becomes too high for direct connection to the machine 
which is to be driven. 

One method of partly overcoming the latter difficulty 
is to operate the turbine at or near the theoretically desir- 
able speed and transmit the power to the driven machine 
through gears which decrease the rotative speed to the 
necessary extent. This method was used with all of the 
early De Laval turbines which were of comparatively small 
capacity. It is now being successfully applied to marine 
propulsion and other purposes for which large units are 
used. It is only a partial remedy in the case of large units, 
however, as the gears necessary for the desired reduction 
and the size of the turbine wheels would both become 
excessive. 



242 



STEAM POWER 




3 

H 



m 

'bJO 



THE STEAM TURBINE 243 

Another and very common method is known as com- 
pounding or staging, . This may be of two varieties. The 
pressure drop in each stage may be limited to that 
which will give a reasonable velocity and a number of 
such stages may be put together in series on one shaft. 
This would give -one set of nozzles and a wheel for each 
stage, the steam discharged from one wheel with very 
low velocity expanding to a lower pressure through the 
nozzles of the next stage and impinging upon the 
wheel of that stage with the resultant high velocity. 
Such an arrangement is known as pressure staging or 
pressure compounding, and is extensively used in large 
turbines. 

The pressure staging method is illustrated in Fig. 160 
as applied to the De Laval type of impulse turbine. The 
combined increase in diameter of wheels and increase 
in length of blades gives the necessary increase in area to 
pass the larger volumes of steam as the drop of pressure 
continues from stage to stage. 

Instead of staging on a pressure basis, staging on a veloc- 
ity basis may be used. In such a case the drop in pressure 
through one set of nozzles is great and the resultant veloc- 
ity high. The steam moving at this high velocity is then 
directed upon the buckets moving at such peripheral velocity 
that they absorb only part of the kinetic energy of the steam, 
discharging it with a lower absolute velocity than that 
with which it entered, but one which is too high to be 
thrown away. The steam then passes through a set of 
stationary vanes which direct it upon the blades of a second 
wheel, in passing through which it gives up still more of 
its kinetic energy with a corresponding further decrease 
of velocity. If the velocity still possessed by the steam 
warrants it, a second set of stationary guide vanes and a 
third set of moving buckets can be supplied for further 
reducing it and by carrying this velocity staging through 
a sufficiently great number of stages any initial velocity 



244 



STEAM POWER 



could be absorbed theoretically without the use of wheels 
with high peripheral speeds. Practically, losses due to 




friction, eddying and other sources limit the number of 
velocity stages to two or three. 



THE STEAM TURBINE 



245 




■cdi 



Second Wheel 



i ^ 

o 




Fig. 161.— Early Form of Curtis Turbine. 



246 STEAM POWER 

Velocity staging is combined with pressure staging in 
the Curtis type of turbine. A section through part of an 
early design of vertical turbine of this type as built by the 
General Electric Company is shown in Fig. 161. The 
turbine illustrated had four pressure stages and each pres- 
sure stage had two velocity stages. 

Many varieties of impulse turbines have been developed 
and all of the larger ones employ several wheels and sets 
of nozzles and diaphragms to obtain the necessary staging. 
The same result has been obtained in some of the smaller 
models by discharging the steam from nozzles on to a set 
of buckets which are able to absorb only a fraction of the 
kinetic energy, catching it at discharge and returning it 
for another passage through the buckets, and so on until 
the greatest practical fraction of the kinetic energy has been 
absorbed. 

A vertical section through a large, horizontal turbine of 
the impulse type is given in Fig. 162. Units of this sort 
are built with different numbers of stages depending upon 
both the total pressure drop for which they are designed 
and the thermal efficiency which is desired. It is obvious 
that the number of stages required to give a certain thermal 
efficiency will increase with the total pressure drop (that 
is, the difference between steam pressures at entrance and 
exit respectively) if the peripheral speed permitted is to 
remain the same. Conversely, if other things remain equal, 
increasing the number of stages increases the thermal effi- 
ciency up to the point where increasing losses overtake 
further possible gains. 

Commercially, the number of stages used in any given 
case is determined as a sort of compromise between first 
cost of unit, operating reliability and money value of thermal 
efficiency. Twenty-two stages are about the upper limit 
at the present time and the great majority are built with 
a smaller number. 

The impulse type is built in all sizes between a unit 



THE STEAM TURBINE 



247 




3A{J(XiOUJ9A.OO 



248 



STEAM POWEE 




Fig. 163. 

Elementary Reaction 
Turbine. 



capable of developing a few horse-power and a unit capable 
of developing about 60,000 horse-power. 

106„ The Reaction Type. If high-pressure steam or 
other fluid be forced into a de- 
vice arranged as shown in Fig. 
163 and free to revolve about 
a vertical axis, the jets blowing out 
of the nozzles will cause the mecha- 
nism to revolve in the direction 
indicated by the arrow. This rota- 
tion is said to be due to the reaction 
of the jets, and the mechanism there- 
fore constitutes a simple form of reaction turbine. By 
increasing the number of nozzles 
any amount of steam could be dis- 
charged and therefore any amount 
of work could be obtained. ,, 

This multiplication of nozzles 
can, however, be more conveniently 
accomplished by fastening radial 
vanes to the periphery of a wheel 
as shown in Fig. 164, the space 
between any two vanes constituting a nozzle through which 
the steam can discharge. By mounting such a wheel 
within a casing as shown in Fig. 165 it forms a simple 
reaction turbine. One of the characteristic differences 
between the impulse and the reaction 
types lies in the distribution of pressures. 
In the impulse type the nozzles are 
fastened into a stationary part of the 
turbine and the drop of pressure occurs 
entirely within the nozzles. The wheels 
are therefore immersed in a space in 
which a uniform lower pressure exists. 
In the reaction type, on the other hand, 
the nozzles are carried on the wheel and 





Fig. 165. 



THE STEAM TURBINE 



249 




250 



STEAM POWER 



there must be a higher pressure on one side of the wheel 
than there is on the other. Since there must also be me- 
chanical clearance between the blade tips and the interior 
of the casing, it follows that the reaction type will be 
handicapped by considerable leakage which does not exist 
in the impulse type, excepting as some of the jet may 
" spill " over the ends of the blades in the latter. 

The difference of pressure on the two sides of the wheel 
also causes a tendency toward motion of the wheel along 
the shaft, or of the wheel and shaft, in a direction away 
from the higher pressure. 

Many unsuccessful efforts have been made to design 
efficient reaction turbines, but no pure reaction type has 

yet been commercialized. 
The turbines commonly 
called reaction turbines are 
really combinations of re- 
action and impulse types. 
One example of what is 
commercially called a reac- 
tion turbine is shown in 
Fig. 166. Alternate rings 
(or rows) of stationary and 
movable blades guide the 
steam as it expands from 
the high pressure at one 
end to the low pressure at 
the other. The stationary 
blades project inward from 
the interior surface of the 
stationary casing and the 
movable blades project out- 
ward from the external surface of the cylindrical rotor. 
The rotor blades act like those of an impulse turbine in 
partly reversing the direction of jets of steam which reach 
them with comparatively high velocities, but they also act 




Fig. 167. 



THE STEAM TURBINE 251 

like the movable nozzles of a reaction turbine since the 
steam in passing through them expands and acquires 
kinetic energy, giving a reaction on discharge. The sta- 
tionary blades serve to redirect the steam so that it strikes 
the next set of moving blades at the proper angle and they 
also serve as nozzles in which velocity energy is acquired. 
This is shown diagrammatically in Fig. 167, in which S 
denotes stationary, and M movable blades. 

The Parsons type, illustrated in Fig. 166, may be 
described as a multistage type in which impulse and reaction 
are utilized in conjunction. 

The balance pistons shown in the figure are used to 
balance the end thrust caused by the difference in pressure 
existing on opposite sides of the wheels in the case of reac- 
tion turbines. Each piston is of such a diameter that it 
presents a surface equal to the blade surface acted upon 
by one of the unbalanced pressures, and by connecting 
across as shown in the figure a high degree of balance is 
secured. 

The overload valve is used to admit high-pressure steam 
to the low-pressure blades for carrying excessive overloads. 
The larger area of the passages through these blades per- 
mits an abnormal amount of high-pressure steam to pass, 
thus giving a high load-carrying capacity with decreased 
economy. 

107. Combined Types. The clearance at the ends of 
the stationary and moving blades in the Parsons type of 
turbine permits considerable steam to leak by, as previously 
explained. This clearance must have almost the same 
length (measured from blade tip to opposing metal) in all 
stages in order to insure freedom from rubbing, but it is 
more detrimental in the high-pressure stages than in the low. 
The high-pressure blades are much shorter than the low- 
pressure blades and a leakage length of a certain amount 
is therefore equal to a greater fraction of the total blade 
length. The density of the high-pressure steam is also so 



252 STEAM POWER 

much greater than that of the low-pressure steam that many 
more pounds can leak through an opening of a given size 
in a given time. In discussions of this character, it should not 
be forgotten, however, that leakage area is determined by the 
dimension already referred to multiplied into a circumference 
and that the circumference is much greater at the lower end. 

Because of these and other reasons many manufacturers 
have come to the conclusion that the impulse type is best 
for the high-pressure end of the turbine and the reaction 
type for the low-pressure end. Many such combinations 
have been produced and they are giving very good results. 

108. Steam Consumption of Steam Turbines. It is 
exceedingly difficult to compare the steam consumption 
of turbines and reciprocating engines in a general way. 
Roughly the steam consumption of the better varieties 
of the two types is of the same order for comparable con- 
ditions with the advantage probably slightly in favor of 
reciprocating engines in the smaller sizes and in favor of 
turbines in the larger sizes. 

It has been shown that the steam turbine operates on 
the complete expansion cycle while the reciprocating engine 
operates on a cycle with incomplete expansion. The tur- 
bine therefore has a certain theoretical advantage because 
its cycle is such as to convert into work a greater amount 
of heat per pound of steam used between given upper and 
lower pressures. 

This choice of different cycles rests on a sound founda- 
tion. This can be appreciated best after studying Fig. 168 
which shows the volumes assumed by steam expanded 
adiabatically from an initial pressure of 150 pounds gauge 
and 100° F. superheat. It will be observed that at the 
lower pressures the volume increases very rapidly with a 
small drop of pressure. If an attempt were made to expand 
steam in a reciprocating engine down to exhaust pressure 
the cylinder would have to be increased in size by a very 
large amount in order to accommodate the rapidly increas- 



THE STEAM TURBINE 



253 



























































































































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J 


s s 


J 8 £ S • S o 



1 s 



« 

•§ 

< 

a; 
oi 



W o 

•si 

Ph 



I ^ 



■gqySQ sanoui-aansseaj 



254 STEAM POWER 

ing volume of steam. As previously stated in Section 40, 
the actual amount of additional energy recovered from the 
steam would not be worth enough to balance the increased 
friction loss of the larger parts and the increased invest- 
ment in the larger engine. Further, some pressure differ- 
ence is required to cause the steam to flow through the 
comparatively restricted exhaust ports and valves in the 
short time available and absolutely complete expansion is 
thus really impossible even if it were desirable. 

With the steam turbine, experience has shown that it 
is often economical to expand the steam down to a low back 
pressure and as there are no restricted exhaust ports and 
valves no pressure differential is required. This is of great 
importance as a very small pressure drop at low pressures 
makes available a very large amount of energy in com- 
parison with the result of a corresponding pressure drop 
at higher pressures. 

For example, trial on a T(f> or Mollier Chart will show 
that saturated steam in expanding with constant entropy 
from 200 pounds absolute to 0.5 pound absolute makes 
available almost as much energy in dropping from 15 pounds 
to 0.5 pound as it does in dropping from 200 to 15. Further, 
such trial will show that at the extremely low pressures a 
very small pressure drop liberates a relatively tremendous 
quantity of energy. 

As a result of these characteristics of steam, the turbines 
which are built to give low steam consumption are designed 
to operate with very low back pressures, that is very " high " 
vacuums. In the larger sizes this introduces real diffi- 
culties in design. The entire volume of steam at the lowest 
pressure has to flow through the last set of blades and these 
must be extremely long or carried on a wheel of large diam- 
eter, or both, in order to give the necessary area for passage 
of the steam. Such difficulties have led to numerous 
" double flow " designs in which the steam is expanded from 
initial pressure to some lower pressure in the ordinary way 



THE STEAM TURBINE 



255 



and is then introduced into a section in which it can divide 
and flow both ways through two opposed but similar sets of 
blades. One double-flow arrangement, as applied to an 
impulse turbine, is shown diagrammatically in Fig. 169 




i 



< 



fe 



and another, applied to a reaction turbine, . is shown in 
Fig. 170. 

In some of the larger sizes of turbines the unit has been 
broken up into two or more parts. Each part is a complete 
turbine but is built for a smaller pressure range than that 



256 



STEAM POWER 




THE STEAM TURBINE 257 

through which the steam is expanded by the complete unit. 
The " high pressure unit " receives steam from the boilers 
and expands it down to some intermediate pressure. It 
exhausts into one, or two, " low pressure units " which 
expand the steam down to the lowest pressure and discharge 
it to the condenser. 

This arrangement is similar to compounding as applied 
to reciprocating engines but it has an entirely different 
purpose. Compounding of turbines in this way offers 
greater flexibility of design. The designer can choose dif- 
ferent rotative speeds for the high and low pressure units 
instead of having to use some sort of compromise as is neces- 
sary when all elements are carried on a common shaft. 
Compounding offers another advantage which is of impor- 
tance in the larger sizes. The size and complexity of each 
of the parts is less than if the entire turbine were built in 
a single unit and it is logical to assume that this adds to 
the operating reliability of extremely large units, if all 
other features are alike. 

The very real thermal advantage to be gained by using 
low back pressures with steam turbines is shown by results 
of tests which indicate that lowering the back pressure 
by one inch of mercury will increase the economy by 
from 3 to 10 per cent, depending upon the type of tur- 
bine, the back pressure under consideration and upon 
other factors. 

Superheat is also very effective in improving the 
thermal efficiency of the steam turbine. In general, every 
ten degrees of superheat causes a saving of 1 per cent in 
the weight of steam required for a given output. 

109. Low Pressure Turbines. Experience has shown 
that reciprocating engines are fully the equal of turbines 
in the high pressure ranges, in many cases they are even 
superior. The turbine, on the other hand, has the advan- 
tage at low pressures and in cases where great ratios of 
expansion are used. It was at one time suggested that 



258 STEAM POWER 

these characteristics should be recognized by building 
mixed plants, using reciprocating engines for the first part 
of the expansion and exhausting these into turbines at or 
near atmospheric pressure. 

Under ordinary conditions this is not an economical 
solution as the investment is so high that any thermal 
gain which can be obtained is not sufficient to balance the 
increased capital charges to say nothing of increased com- 
plications. However, the scheme has been used to advan- 
tage in increasing the capacity and thermal efficiency of 
reciprocating plants which were already installed. 

The ability of the turbine to handle low pressure steam 
to advantage has given rise to the use of low pressure 
turbines in many different ways. As examples, the ex- 
haust of hoisting engines, steam hammers, and other 
apparatus commonly exhausting at atmospheric pressure 
is now frequently led to one or more low pressure tur- 
bines in which it is expanded with the recovery of a large 
amount of power from what would otherwise be waste 
steam. 

When low pressure turbines are used in this way it fre- 
quently happens that the demand for steam on the part 
of the turbine and the make of steam on the part of the 
primary user are so different from instant to instant that 
some device must be used to store steam between the two. 
The device used for this purpose is known as a regenerator. 
It consists of some sort of closed vessel in which steam can 
be mixed with and condensed in hot water. When the make 
exceeds the demand of the turbine the pressure and temper- 
ature within the regenerator rise; when the demand of the 
turbine exceeds the make the pressure and temperature 
within the regenerator fall. 

110. Steam Turbo-generators. The steam turbine is 
ideally suited to the driving of electric generators of the 
alternating current type as the desirable speeds for the two 
devices fall in the same general range of values. This fact, 



STEAM TUEBINE 



259 



coupled with the ease with which such units can be con- 
structed in large capacities, the comparatively low cost 
and the high thermal efficiency attainable, has resulted in 



3200 



J100 



82000 




I 




II': 

STEAM TURBINE PROGRESS 
















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24000 
P. 

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60000 



50000 



40000 •- 

4) 

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30000 s 



20000 



10000 



1890 



1900 



Year 



1910 



1920 



Fig. 171.— Curves Showing Progress of tha Steam Turbine. 

bringing the steam turbine into almost universal use for 
this purpose where steam is used as the means of producing 
power. 

The history of the development of the large Turbo- 
generator is shown in Fig. 171. 



PROBLEMS 

1. A steam turbine produces one horse-power hour at its shaft 
for every 30 lbs. of steam supplied. The initial pressure is 200 
lbs. absolute and the steam is superheated 200° F. The turbine 
exhausts against a back pressure of 14 lbs. absolute. 

Find the thermal efficiency on the assumption that heat of 
liquid at exhaust temperature is not chargeable to the turbine. 

2. Develop a complete expansion cycle for one pound of mate- 
rial used under the conditions of Prob. 1 and find the energy 



260 STEAM POWER 

made available per cycle. From this value determine the number 
of pounds of material theoretically required per horse-power hour 
and compare with the value given in Prob. 1. 

3. Find the additional quantity of energy which would theoret- 
ically be made available per pound of steam in above problems if 
the back pressure could be lowered to \ lb. absolute. 

4. Develop a complete expansion cycle from an initial pressure 
of 225 lbs. absolute with a superheat of 200° F. to a back pressure 
of | lb. absolute. Assume that this is to be divided up into six 
parts, each making available the same quantity of energy. Find 
the pressure drop for each part. Note that this is most easily 
done with the help of the Mollier chart. 

5. A steam turbine receives steam at a pressure of 225 lbs. 
per square inch absolute and with a superheat of 190° F. and 
exhausts into a condenser in which a pressure of f lb. per square 
inch absolute is maintained. The turbine is direct connected to 
an electric generator and produces a K.W.-hour on 12 lbs. of steam. 
If a K.W.-hour is equivalent to 3411 B.t.u., what is the thermal 
efficiency of the combination? 

6. Develop a complete expansion cycle for the conditions of 
Prob. 5 and determine the pounds of steam which would be re- 
quired theoretically to develop energy equivalent to 1 K.W.-hour. 
Compare with the value given in Prob. 5. 

7. Determine the velocity theoretically attainable by expanding 
steam in one step from the initial to the final conditions of Prob. 
5 above. What would be the value of the kinetic energy of such 
a jet per pound of steam flowing? 

8. Determine the shape of a nozzle required to discharge 
1000 lbs. of steam per hour, initial conditions being 100 lbs. per 
square inch absolute, and dry saturated steam; final pressure 
being 2 lbs. absolute. 

9. Determine velocity and kinetic energy of jet in Prob. 8. 



CHAPTER XIV 
CONDENSERS AND RELATED APPARATUS 



111. The Advantage of Condensing. The lowest pres- 
sure to which an engine or turbine can expand steam, that is 
the exhaust pressure, is determined by the pressure prevail- 
ing in the space into which the steam is exhausted. With a 
given initial pressure the amount of work which can be ob- 




Fig. 172. 

tained theoretically from a given weight of steam increases 
as the exhaust or back pressure decreases, as shown by the 
areas of the two diagrams in Fig. 172, and experience has 
shown that at least a part of this theoretical increase can be 
obtained in real engines. It is therefore desirable to ex- 
haust into a space in which the lowest possible pressure 
exists when the work obtained per pound of steam is the 
only consideration. 

The most available space into which an engine can 

261 



262 STEAM POWER 

exhaust is that surrounding the earth and already occupied 
by the earth's atmosphere. The pressure in this space 
is approximately equal to 14.7 lbs. per square inch at sea 
level and is due to the weight of the atmosphere. Since 
the superincumbent column of atmosphere decreases in 
depth as one moves upward, its weight also decreases and 
atmospheric pressure therefore averages less than 14.7 
lbs. per square inch at high altitudes and has a greater 
average value at points below sea level. 

If it is desired to exhaust into a pressure lower than 
atmospheric a means of maintaining such an abnormal 
pressure within some sort of vessel must be devised. It 
is the purpose of a condenser and its associated apparatus 
to make available a space in which such a low pressure can 
be maintained. Its method of operation will be considered 
in later sections. 

There is also another advantage which may be ob- 
tained by the use of a condenser. It often happens that 
the water available is not well adapted to use in boilers. 
It may be salt water as in marine practice, or it may contain 
a number of undesirable gases and solids in solution as often 
occurs in stationary practice. Some types of condensing 
apparatus are so arranged that the steam exhausted from the 
engine is converted into liquid water without admixture 
and can therefore be returned to the boiler as practically 
pure water, thus largely eliminating the troubles that 
would ensue from the use of poor feed water. 

112. Measurement of Vacuum. Assume that some non- 
volatile liquid, that is, a liquid that did not vaporize, 
could be found and also that it contained no gases in solu- 
tion. If a long tube were inserted in a vessel filled with 
such a liquid and had its upper end connected with some 
form of vacuum pump which could remove air from its 
interior, as shown in Fig. 173, liquid would rise in the tube 
as the air was removed. Removal of air would result in 
lowering the pressure within the tube, but the constant 



CONDENSERS AND RELATED APPARATUS 263 



^ 



M 



.-Pn 



atmospheric pressure on the liquid surface outside the tube 

would then force liquid up the tube to such a height that 

the pressure p a of air still in the tube 

plus the pressure due t«o the column 

of liquid of height h within the 

tube just equaled the pressure due 

to the atmosphere on the surface 

of the liquid in the vessel. If the 

pump could remove all of the air 

from the tube, liquid would rise to 

such a height that the pressure 

exerted by it on a plane passing 

through the lower surface just 

equaled that of the external atmos- 
phere. 

The same result could be at- 
tained by using a tube closed at 

one end, rilling it with the liquid, 

and then inverting so that the end 

rested in the liquid as shown in 

Fig. 174. If the tube were long 

enough, the liquid would drop to some such 
point as shown, under which conditions the 
height of liquid would just balance atmos- 
pheric pressure. This would only be true if 
the liquid did not volatilize and did not contain 
gases in solution; with these assumptions the 
space above the liquid in the tube would con- 
tain absolutely nothing. This space would be 
said to be perfectly vacuous, or a perfect vacuum 
would be said to exist in that part of the tube. 
A device of this character is used to 
measure the pressure of the atmosphere and 
is known as a barometer. Mercury is used 

as the liquid because its high density makes it possible 

to use a short tube and because it may be considered 



vzszzzszzziimA 



Fig. 173. 



L 



Fig. 174. 



264 



STEAM POWER 



as non-volatile at ordinary temperatures. The average 
atmospheric pressure at sea level, equal to 14.7 lbs. per 
square inch approximately, can support about 30 ins. 
of mercury, so that this figure is generally taken as the 
standard sea level barometer reading. An atmospheric 
pressure of less than 14.7 lbs. would be shown by a barom- 
eter reading of less than 30 ins.; a greater atmospheric 
pressure by more than 30 ins. Corresponding values of 
atmospheric pressure and barometer reading are given 
in Table VII. To this have also been added the altitudes 
to which the different values would correspond if a pressure 
of 14.7 lbs. existed at sea level and there were no variations 
of atmospheric pressure excepting those due to change of 
elevation. Values of this type can only be roughly approx- 
imate, because local barometric variations are constantly 
occurring and the sea-level atmospheric pressure varies 
both sides of 14.7 lbs. 

TABLE VII 
Atmospheric Pressure, Barometer Reading and Altitude 

(Negative signs mean distance below sea level.) 



Barometer, 


Atmospheric Pressure, 


Altitude, 


Inches of Mercury. 


Pounds per Square Inch. 


Feet (Approximate). 


25.00 


12.27 


4750 


26.00 


12.76 




26.50 


13.01 


3250 


27.00 


13.25 




27.50 


13.49 


2250 


28.00 


13.74 




28.50 


13.98 


1300 


29.00 


14.23 




29.25 


14.35 




29.50 


14.47 


450 


29.75 


14.60 




30.00 


14.72 


Sea level 


30.25 


14.84 




30.50 


14.96 


-450 


30.75 


15.09 




31.00 


15.21 


-900 



CONDENSERS AND RELATED APPARATUS 265 

The exact value of standard atmospheric pressure 
at sea level is taken at 29.921 ins. of mercury, which is 
equal to 14.696 lbs. per square inch and corresponds to 
the 760 mm. of mercury, used by scientists as standard. 

A tube with both ends open and arranged as shown 
in Fig. 173 can be used to measure the degree of vacuum 
existing in the space to which its upper end is connected, 
and many vacuum gauges are constructed on this principle, 
using mercury as the liquid. The extent to which the 
pressure is lowered in the top of the tube is indicated by 
the height to which the mercury column rises and this 
height in inches is used as a measure of the vacuum. Thus 
if a perfect vacuum were created and if the atmospheric 
pressure were equal to 14.7 lbs. the gauge would show 
about 30 ins. of mercury above the level in the reservoir. 
If the vacuum were less perfect the gauge would show a 
shorter column. 

It should be noted that the reading of the vacuum 
gauge does not give the pressure existing in the vacuous 
space, but gives the amount by which the pressure has been 
reduced below that of the atmosphere, the difference 
being expressed in inches of mercury. By subtracting this 
reading from the existing atmospheric pressure expressed 
in the same units, the absolute pressure in the partially 
vacuous space (expressed in inches of mercury) is obtained. 

It is obvious, therefore, that a vacuum-gauge reading 
of say 28 ins. of mercury does not always mean the same 
absolute pressure. With a barometer reading of 28 ins. 
it would represent a perfect vacuum; with a barometer 
reading of 30 ins. it would represent a partial vacuum, the 
absolute pressure in the partially vacuous space being 
equal to 2 ins. of mercury. 

113. Conversion of Readings from Inches of Mercury 
to Pounds per Square Inch. It is often necessary to con- 
vert readings of pressure in inches of mercury into pounds 
per square inch, This can be done with sufficient accuracy 



266 STEAM POWER 

under ordinary circumstances by multiplying the inches 
of mercury by the constant 0.4908. Thus, 

Barometer in inches X 0.4908 = atmospheric 

pressure in pounds per square inch . . . (74) 
and 

(Barometer in inches — vacuum gauge in inches) 
X 0.4908 = absolute pressure in partially 
vacuous space in pounds per square inch. . (75) 

ILLUSTRATIVE PROBLEM 

A vacuum gauge constructed like that shown in Fig. 167 
reads 27 ins. when the barometer reads 29.5 ins. What is the 
absolute pressure in the partial vacuum above the mercury? 

The absolute pressure is equal to 29.5—27=2.5 ins. of mer- 
cury, which is equal to 

2.5X0.4908 = 1.227 lbs. per square inch. 

114. Principle of the Condenser. A perfect vacuum 
could be created in any closed vessel with impenetrable 
walls if a pump could be devised which could remove all 
material contained within that vessel. Or, any degree of 
vacuum can be maintained in any partially closed vessel 
by fitting to it a pump which can remove all material 
flowing into the vessel as fast or faster than it enters, raise 
the pressure of this material to atmospheric or higher and 
discharge it. 

The latter principle is made use of in real condensers, 
a pump of some form, or an equivalent, removing from the 
condenser the material exhausted by the engine and inleak- 
age from the atmosphere, and discharging it at atmos- 
pheric pressure at a sufficiently rapid rate to maintain 
the desired vacuum. If the condenser and connections 
could be made leak proof, the pump or equivalent would 
have to handle only the material exhausted from the engine. 

A steam engine exhausts a mixture of steam, water 
and gases, the gases being a mixture of those originally 



CONDENSERS AND BELATED APPARATUS 267 

dissolved in the boiler-feed water and air which leaks into 
those parts of the system in which a partial vacuum is main- 
tained. If the pump had to handle the same volume of 
material as is exhausted by the engine, no gain of work 
would result from condensing, because the pump would 
have to do at least as much work in raising the pressure 
of this material to atmospheric and discharging it as could 
be obtained by allowing it to expand in the engine. 

Steam, however, occupies a much larger volume than 
water at the same temperature and pressure. Thus steam 
at 212° F. occupies a volume of about 26.79 cu.ft. per 
pound, but water at the same temperature and pressure 
occupies a volume of only about 0.0167 cu.ft. per pound; 
at a temperature of 120° F. which is often used in condensers, 
the specific volume of steam is about 203 and that of water 
only 0.0162. Therefore, if the steam is condensed after 
exhaust from the engine and before entering the pump to 
be discharged to atmosphere, the pump work is greatly 
reduced. The volume of the condensate is almost negli- 
gible in comparison with the volume of steam exhausted, 
and the work of pumping it is almost negligible in compari- 
son with the work it made available in the engine. 

Gases contained in the exhaust steam cannot be lique- 
fied and must be pumped as gases. The work required 
to pump them can, however, be reduced by lowering their 
temperature as far as possible. 

The condenser equipment may be regarded as con- 
sisting of a combination of a partially closed vessel and 
some form of pump. The vessel is so constructed that a 
low temperature can be maintained within it and that 
large quantities of heat can be removed from it for the 
purpose of condensing the exhaust steam and of cooling the 
contained gases. This is generally done by using large 
quantities of cool water. 

The absolute pressure within the condenser is made 
up of two parts. The two parts are, (a) that due to the 



268 STEAM POWER 

water vapor, since the space over the condensed water will 
always be filled with saturated steam at the same tempera- 
ture (approximately) as that of the water, and (6) that due 
to any gases present. 

The pressure of the saturated steam (water vapor) 
can be found from the steam tables opposite the temperature 
existing in the condenser and it is the pressure which would 
exist in the condenser of an ideal system in which no gases 
were mixed with the working substance. The pressure 
of the gases can be found by subtracting from the total 
measured pressure in the condenser the pressure exerted 
by the water vapor as shown in the steam tables. The 
pressures exerted by the water vapor and gases are spoken 
of as partial pressures, since their sum makes up the total 
pressure within the condenser. 

The presence of gases causes a two-fold loss. First, 
it increases the pressure against which the engine has to 
exhaust, thus raising the back-pressure line on the diagram 
and decreasing the work area. Second, it increases the 
work which must be done by the pump which otherwise 
would only pump the condensate and such saturated water 
vapor as accompanied it. 

115. Types of Condensers. The condensers actually 
used in steam plants can be roughly divided into two types, 
as 

(a) Contact condensers and 

(6) Non-contact condensers. 
In the first type the water which is used for condensing 
and cooling is intimately mixed with the exhaust from the 
engine within the condensing vessel, and the resultant 
mixture of condensing water, condensate and gases is drawn 
out of this vessel and discharged to atmosphere by the 
pump. 

In the second type condensing water flows on one side 
of metal surfaces of some sort and the exhaust is led over 
the other side, the heat being transmitted through the 



CONDENSERS AND RELATED APPARATUS 269 



Condensing 
Water In 



from Engine 



metal. In condensers of this type the condensate and 
gases are not mixed with the condensing water and the 
condensate can therefore be returned to the boiler as feed 
water with the advantages already mentioned. 

116. The Jet Condenser. One of the earliest forms of 
contact condensers which is still very widely used for 
moderate vacuums is commonly 
known as the jet condenser. The 
principle of operation of the jet 
condenser is shown in Fig. 175. 
Water, under pressure, entering 
as indicated, is broken up into 
fine streams or jets and sprayed 
into the exhaust coming from 
the engine. The resultant mix- 
ture flows downward into the 
neck of the condensing vessel or 
" condenser head " and is re- 
moved by some form of pump. 
This pump handles gases, vapors 
and water and is known as 
a vacuum pump, a wet-vacuum 
pump, or a wet-air pump, the 
term wet signifying that it han- 
dles the water as well as the 
gases. 

The pressure within such a 
condenser head would be theo- 
retically equal to that corresponding to the temperature 
of the resultant mixture if no gases were present. In 
practice the pressure of the water vapor would roughly 
correspond to the average temperature near the top of the 
vessel and there would be a partial pressure due to gas 
as well. This gas would consist of that brought over by 
the engine exhaust plus that released from the cooling or "cir- 
culating " water under the low pressure within the condenser. 




^Mixture to Pump 

Fig. 175. — Jet Condenser. 



270 



STEAM POWER 



Details of a, complete jet condenser and of the method 
of connecting it to an engine are given in Fig. 176. The 
atmospheric relief valve is installed in all condensing sys- 
tems and is arranged to open automatically and thus dis- 
charge exhaust steam to atmosphere if the pressure within 
the system rises to atmospheric, that is, if the " vacuum 
is lost." 




Fig. 176— Jet Condenser and Method of Connecting to Engine. 



With the jet condenser the pressure might start to 
rise because of slow action or even stoppage of the pump. 
As the condenser head filled up the rising water would 
ultimately entirely cover the jet and condensation would 
then practically cease. In the arrangement shown in 
Fig. 176 there is an additional safety device which breaks 
the vacuum in the exhaust system if the water in the head 



CONDENSERS AND RELATED APPARATUS 271 

rises above a certain height, thus preventing the external 
atmospheric pressure from forcing this water back along 
the exhaust pipe and into the cylinder, an event which 
would probably result in a wrecked engine. 

The jet condenser here described is known as a parallel- 
flow type, because everything within the condensing vessel 
flows in the same direction. The gases and vapors handled 
by the pump theoretically have the same temperature 
as that of the mixture with which they flow out at the 
bottom of the condenser head. The temperature of this 
mixture therefore determines the temperature of the gases 
and vapors pumped. 

There are numerous forms of contact condensers which 
more or less closely resemble the types of jet condenser 
just described. They are properly all classed as jet con- 
densers, but more often are given distinguishing names. 

One very common form of contact condenser is generally 
known as a barometric condenser. It consists essentially 
of a condenser head, similar to that used with the jet con- 
denser already described, and a tail pipe or barometric 
pipe which partly or wholly takes the place of the wet- 
vacuum pump by removing part or all of the mixture formed 
within the condenser. One model of such a condenser is 
shown in Figs. 177 and 178. 

The exhaust from the engine enters the head through 
the large pipe shown and divides into two parts, one part 
passing down through the center of the head and the re- 
mainder flowing downward in the annular space A. The 
circulating or injection water enters as shown, is divided 
by the spraying cone and injected into that part of the 
exhaust, which enters the central tube of the condenser. 
The mixture thus formed flows downward and finally 
meets the discharge from the lower end of the annular 
space A, which is then condensed. The mixture of injec- 
tion and condensing water together with such gases as have 
been entrapped, then flows downward into the tail pipe, 



272 



STEAM POWER 



which is over 34 ft. in length and which dips into the " hot 
well " at its lower end. As atmospheric pressure can only 



/To Vacuum or 
Dry Air Pump 



Air Cooler 



Exhaust 



Drain Injection 

I Water 




WateT for Cooling 
"Air" 



Fig. 177. — Barometric Condenser. 



support a column of water about 34 ft. high, the tail pipe 
forms an automatic wet-vacuum pump, water flowing from 
it as rapidly as it accumulates within it. 



CONDENSERS AND RELATED APPARATUS 273 




274 



STEAM POWER 



^e- Atmospheric Relief Valve 



Experience has shown that the maintenance of a high 
vacuum with this type of condenser depends upon the ex- 
tent to which gases are removed from the condenser head. 
These gases are generally called air, as the greater part of 
them is air. In the type illustrated such " air " as is 
not trapped by the descending mixture rises through the 

hollow spraying cone, then 
through the air cooler and 
flows out through the pipe 
indicated to the vacuum or 
dry-air pump. The air in 
rising through the center 
of the spraying cone is 
cooled by the water flowing 
around it, and it is further 
cooled by coming into con- 
tact with water as it works 
its way through the air 
cooler. This results not 
only in lowering its tem- 
perature, but also in caus- 
ing the condensation of a 
great deal of the water 
vapor accompanying it. 
This condensed vapor col- 
lects in the space surround- 
ing the air cooler and 
flows down into the head 
through the drain shown. The vacuum pump, therefore, 
handles cool gases containing little water vapor and prac- 
tically no liquid water. It is sometimes called a dry-air 
pump or dry-vacuum pump for this reason. 

The entrainer shown in the exhaust system in Fig. 178 
is so shaped that water collecting in the exhaust piping 
and flowing into the entrainer is picked up by the exhaust 
steam and carried into the condenser. 




Fig. 179. — Baragwanath Barometric 
Condenser. 



CONDENSERS AND RELATED APPARATUS 275 

The flow of steam and injection water in this condenser 
is parallel, but the material on its way to the dry- vacuum 
pump flows upward and the cooling water flows downward 
so that counter-current flow is used in this part of the appa- 



SYPHONING ITS 
WATER FROM 
£ TANK OR FLUME 



BARAGWANATH 

CONDENSER 

ORDINARY 

SETTING 




Fig. 180. 



ratus. This has the advantage of bringing the air leaving 
the condenser into contact with the cooling water just as 
it enters and therefore when it has its lowest temperature. 

A somewhat similar condenser, arranged so that it 
requires no pump, is shown in Figs. 173 and 174 (a) and (6). 



276 



STEAM POWER 



Exhaust and injection water mix as shown, the quantity 
of injection water being regulated by the hand wheel on 
top of the condenser. The mixture flows downward through 
the narrow neck and the velocity attained in this part of the 
tail pipe is so high that all air and similar gases are swept 
along with the current. 

Fig. 180 (6) shows an 
arrangement not requir- 
ing a circulating pump. 
For starting, the valve 
V is opened, allowing 
water to flow into the 
lower part of the tail 
pipe. This creats a par- 
tial vacuum, and atmos- 
pheric pressure then 
forces water up the in- 
jection pipe and into 
the condenser head. The 
valve V is then closed 
and the condenser con- 
tinues to siphon its own 
water. Because of this 
action this type is often 
called a siphon conden- 
ser. By supplying a 
circulating pump as in- 
dicated in Fig. 180 (a) 
it can be converted into a barometric condenser similar to 
the type already discussed except for the fact that it requires 
no air pump. 

The barometric or tail pipe of any barometric condenser 
can be replaced by any kind of a pump, and centrifugal 
pumps are often used for this purpore. Centrifugal air pumps 
have also been devised and are in use. 




Fig. 181.- 



-Westinghouse-Leblanc Air 
Pump. 



CONDENSERS AND RELATED APPARATUS 277 

A " low level jet condenser " in which the barometric 
tube is replaced by a centrifugal pump and in which a sepa- 
rate air pump of a rotary type is used is illustrated in Figs. 
181, 182 and 183. It consists essentially of the condensing 
head and well, combined with a centrifugal tail pump and 
a rotary air or vacuum pump as indicated in Fig. 183. 




Discharge to 
Hot Well 



Water and 
Vacuum. Pump JJ=L , 

Check Valve 



1 



7^ 



Inlet to Well- 



Gate Valve. 




Basement Floor lane* 



V t Submerged not^ 
n I less than 3. £t» 



*&&&&?&*■&£&&*&$■$&& 



Fig. 182. — Westinghouse-Leblanc Condenser. 



Injection water entering through nozzles in the head meets 
the exhaust, and the resultant mixture flows down into 
the well through the large nozzle shown. The liquid is 
continuously removed from the bottom of this well by the 
centrifugal tail pump and discharged to the hot well. The 
air and associated vapors are drawn down the air pipe and 



278 



STEAM POWER 



Exhaust 



discharged by means of the device shown in Fig. 181. 
Water enters the central part of this pump as indicated 
in Fig. 182 and is discharged through the stationary nozzles 
and the moving vanes V shown in Fig. 181. The water 
is thus caused to form a series of " pistons " which move 
rapidly downward in the discharge nozzle N' and which 
trap small plugs or lamina of " air " between them and 

thus discharge the " air " to 
the atmosphere. The con- 
nection marked P is used 
for priming at starting when 
necessary. 

In small units the cen- 
trifugal tail pump may be 
omitted and the design so 
remodeled that all the injec- 
tion water passes through 
the rotary air pump which 
discharges the entire mix- 
ture from the condenser just 
as it discharges the air and 
associated vapors in the 
larger sizes. 

117. Non-contact Con- 
densers. The type called 
the surface condenser is the 
best-known example of non- 
contact condenser. It con- 





Fig. 183. 



-Westinghouse-Leblanc 
Condenser. 



sists essentially of a long vessel of cylindrical, rectangular 
or other section into which the exhaust is discharged and 
through which pass numerous brass or alloy tubes which 
carry the circulating water, and the surface of which acts 
as the condensing and cooling surface. 

One form of surface condenser mounted above the 
pumps which serve it is shown in Fig. 184. The exhaust 
enters at the top of the rectangular shell and works its 



CONDENSERS AND RELATED APPARATUS 279 




o 

U 

a> 
o 

3 
03 

"3 



280 STEAM POWER 

way down over the water-cooled tubes. The condensate, 
mixed with gases and vapors, is drawn from the bottom of 
the shell by the wet-vacuum pump and discharged to the 
hot well. 

The condensing water is forced through the tubes 
of the condenser by means of the reciprocating circulating 
pump, entering the lower tubes at the right-hand end in 
the figure, making two " passes " through the condenser and 
leaving at the top. Because of the path of the water a 
condenser of this type is sometimes called a two-pass or 
double-flow condenser. 

With the arrangement illustrated, the steam which 
condenses upon the upper tubes falls as a rain from tube 
to tube until it finally settles at the bottom and is drawn 
off. The outer surfaces of the lower tubes are therefore 
practically covered with water and this has two disad- 
vantages. First, these tubes carry the coolest circulating 
water and they therefore cool the condensate coming in 
contact with them while the water flowing through them 
is unnecessarily heated. Cooling of the condensate means 
a lower hot-well temperature than would otherwise be 
obtained, but if the condensate is to be used for boiler 
feed, the temperature of water in the hot well should be 
maintained as high as possible, since this water will eventually 
have to be heated to boiler temperature with a correspond- 
ing expenditure of heat. Second, tubes which are being 
used to cool water covering them are of little use as condens- 
ing surface, and hence such surface in a condenser is com- 
paratively inactive. 

The ideal arrangement would carry away the liquid 
condensate as fast as formed, leaving the tubes first entered 
by the condensing water to act as the final condensing 
and cooling surfaces, thus bringing gases and non-condens- 
ible vapors into contact with the coolest surfaces just before 
entering the vacuum pump. Numerous designs which 
approximate this ideal have been developed recently and 



CONDENSERS AND RELATED APPARATUS 281 

they give better results than do the earlier and simpler 
forms. The improvement is shown by the values of con- 
densing surface per developed horse-power of engine. In 
early designs it was customary to supply 2 \ sq. ft. of tube 
surface or more per horse-power. Some of the most recent 
installations are giving better vacuums with only 1 sq. ft. 
per horse-power. 

One of these condensers passes tne condensate through 
a set of tubes so located that the engine exhaust strikes 
them before impinging on any tubes carrying condensing 
water. This results in a partial condensation of the exhaust 
and raises the temperature of the condensate within the 
tubes to very near that of the exhaust, thus heating the 
boiler feed to a temperature practically corresponding to 
the exhaust temperature of the engine. 

Surface condensers are commonly operated with a 
vacuum of from 24 to 26 ins. of mercury when used with 
reciprocating engines and with a vacuum of 28 to 29 ins. 
when receiving the exhaust of steam turbines. When 
operated at the lower vacuums wet-vacuum pumps are 
generally used, but the best types of dry-air pumps must 
be installed in combination with well-designed condensers 
when the higher vacuums are sought. An installation of 
a surface condenser and necessary auxiliaries as applied to 
a steam turbine of moderate size is shown in Fig. 185. 
The steam exhausted by the turbine enters the top of the 
condenser shell and spreads out over the tubes. As it is 
condensed it gravitates to the lower part of the shell and 
finally flows into the " hot well " attached to the lower 
part of that shell. It is removed from the hot well by the 
" hot well pump " which discharges it to storage tanks or 
heaters, depending on the layout of the plant. 

The cooling water is supplied by the " circulating pump " 
shown. After making two " passes " through the tubes, 
it flows away at the " circulating water overflow." 

The noncondensible gases are drawn from an " air box " 



282 



STEAM POWER 



near the bottom of the condenser by means of an hydraulic 
air pump which will be described later. They flow through 
the pipe indicated as " dry air suction/' enter the " hydrau- 
lic vacuum pump," and are discharged (in intimate mixture 
with water) into the " sealing tank." Here the noncon- 
densible gases separate and escape through the " sealing 
tank vent." The water required by the hydraulic vacuum 
pump is circulated by the " hydraulic supply pump " and 



Turbo-Generator 



Atmosphere 




Supply Water fckalinf; Tan 

Wake-up Overflow 



Fig. 185. — Surface Condenser and Necessary Auxiliaries as Applied to 
Steam Turbine of Moderate Size. 



any loss is made good through the line marked " supply 
water make up." 

In the case shown, the hot well pump and the hydraulic 
supply pump are driven by a single steam turbine, the 
exhaust of which is carried to the feed water heater and 
heats the condensate on its way to the boilers. 

118. Vacuum Pumps or Air Pumps. In an earlier sec- 
tion, attention was called to the fact that some means must 
be provided for removing noncondensible gases, or " air," 



CONDENSERS AND RELATED APPARATUS 283 

from condensers. This may be done by a pump which 
handles both the condensate and the noncondensible mate- 
rial. Such a pump is known as a " wet vacuum pump." 
A section of a pump of this kind is shown in Fig. 176 and 
another in Fig. 184. 

For larger installations, and particularly those in which 
a very high vacuum is desired, it is customary to use one 
pump for handling the condensate and another pump for 
the noncondensible gases. The latter is known as a " dry 
vacuum pump " or more commonly, an " air pump." 

The earliest dry vacuum pumps were merely recipro- 
cating air compressors which worked between a high vacuum 
and atmospheric pressure. That is, they received " air " 
at the pressure existing in the condenser and compressed 
it to such a pressure that they could discharge it to atmos- 
phere. Such pumps are still in use. They are commonly 
arranged with crank and fly wheel and are known as " Rota- 
tive Dry Vacuum Pumps " or " R. D. V. Pumps." 

The air cylinder of such a pump is shown in Fig. 186. 
Air enters at the flange indicated and flows into the right- 
hand end of the cylinder through the upper set of valves 
as the piston moves to the left. On the return stroke the 
air is discharged without compression through the lower 
set of valves, and travels through a passage outside the 
cylinder to the space being opened up on the other side 
of the piston, entering through the valves shown at the left 
and above the bottom of the cylinder. When the piston 
makes its next stroke to the left, the air entrapped in the 
left-hand end of the cylinder is compressed and discharged 
through the valves at the bottom. Such a pump is de- 
scribed as " two stages in one cylinder." The right-hand 
end merely serves as a " loader " for the left-hand end 
and all compression occurs in the latter. This " loading " 
or " two stage " feature is introduced for the purpose of 
increasing the volumetric efficiency, that is, the amount 
of air handled per revolution. 



284 



STEAM POWER 



With the arrangement shown, the right-hand end of the 
cylinder and its clearance are completely filled with air at 
suction pressure at each stroke. That part filling the piston 
displacement is air drawn from the condenser and that part 
filling the clearance is air left over from the previous return 
stroke. All of the air filling the piston displacement is 



Water Jacket 







Discharge Valves ufy/y/A 
Second Stage <S&^ 



Discharge Valvea 
1st Stage 



Discharge, Compressed Air 

Fig. 186. — Two-stage Single Air Cylinder. 



transferred to the other end for compression and discharge 
during the return stroke of the piston. 

If the pump were designed to compress to atmospheric 
pressure at the right-hand end and to discharge the air to 
the atmosphere, its clearance would be filled with air at 
atmospheric pressure at the end of the discharge. When 
the piston moved to the left, this clearance air would have 
to expand to condenser pressure before any air could enter 
through the suction valves from the condenser. Experience 
shows that a relatively large part of the total piston dis- 
placement is uselessly wasted in such re-expansion of air 
caught in the clearance. For this reason, practically all 



CONDENSERS AND RELATED APPARATUS 285 



the better rotative dry vacuum pumps embody in their 
design some feature intended to minimize this effect. 

After reciprocating air pumps had been developed to 
a point where little further improvement seemed possible, 




u 






€ 


-a aia \ 


.a e 

aw 


t 





two radically different types of air pumps were introduced. 
One kind may be called the " hydraulic type " and the other 
the " steam jet type." 

One example of hydraulic air pump is shown in Figs. 



286 



STEAM POWER 



181, 182 and 183. Another, as built by the Worthington 
Pump & Machinery Corporation, is illustrated in Fig. 185 
and is shown in detail in Fig. 187. 



gs£EEL_^t" > 



Air Suction 




Fig. 1876. — Hydraulic Vacuum Pump. 



In the type shown in Fig. 187, water supplied by a 
centrifugal pump enters the air pump at the hurling water 
inlet. It flows out of the head of the air pump through 
the annular nozzle indicated and enters the revolving jet 
wheel with a high velocity. The water leaves the wheel 



CONDENSERS AND RELATED APPARATUS 287 

in the form of a number of jets of approximately rectangular 
section, each jet traveling a helical path leading it into the 
throat of the discharge tube. 

In traveling across the space between the revolving 
wheel and the throat, the jets entrap air between them and 
carry this air into the discharge tube. In passing through 
the lower part of the discharge tube, the velocity of the 
mixture decreases and the pressure increases so that the 
mixture can be discharged into the sealing tank against 
a small head of water plus atmospheric pressure on the sur- 
face of the water. 

In the tank the air separates and escapes through the 
vent. The water is used over and over, any loss by evapo- 
ration or spillage being made up as convenient. 

In the steam jet type, one or more steam jets entangle 
or entrain the condensible gases while moving at a high 
velocity and the mixture then passes through an expanding 
tube or nozzle in which its velocity is reduced with a cor- 
responding increase of pressure. The design is such that 
the pressure rises to a value sufficiently above atmospheric 
to make possible the discharge of the mixture to the atmos- 
phere. 

Such a device is shown diagrammatically in its simplest 
form in Fig. 188. Steam at a high pressure enters the 
expanding steam nozzle designed to discharge it at a high 
velocity and with a pressure slightly lower than that carried 
in the condenser. The jet of steam passes through the space 
indicated as " entraining space " and into the " discharge 
tube." While passing through the entraining space, the 
jet picks up or entrains some of the noncondensible material 
held in that space and the mixture enters the discharge 
tube. The taper or flare of the discharge tube below the 
" neck " is so proportioned that the high velocity of the 
mixture is reduced and the pressure correspondingly in- 
creased to a value equal to or greater than atmospheric. 

The entraining space is connected to that part of the 



288 



STEAM POWER 



High Pressure 
Steam Inlet 



Steam Nozzle 
■^fAir Inlet or 
Suction Flange 



hroat or Neck 



Diffusor or 
Discharge Tube 



condenser in which the nonconclensible gases collect. As 
this material is constantly removed from the entraining 

space by the steam jet, a 
continuous flow from con- 
denser to air pump results. 

Experience has shown 
that a single stage pump 
like that illustrated in Fig. 

188 is not the best possible 
arrangement. As a con- 
sequence, two and three 
stage pumps have been pro- 
duced, the two stage being 
the type now generally used. 
Such a pump is illustrated 
in Fig. 189, and in Fig. 190 
is shown one way in which 
it is connected into a con- 
denser installation. 

The device shown in Fig. 

189 is called the Radojet 
Vacuum Pump, taking its 
name from the peculiar ar- 
rangement at the second 
stage which is described 

later. Instead of using one steam nozzle leading steam into 
the entraining space as shown in Fig. 188, a group of small 
nozzles is used. This is done for the purpose of breaking 
the steam up into a number of small jets, thus increasing 
the amount of air entrained by a given weight of steam. 

The mixture discharged from the diffusor of the first 
stage enters the second entrainment space. Steam admitted 
at the second stage steam inlet expands through the pecul- 
iarly shaped nozzle indicated and is discharged at high 
velocity in the form of a circular sheet. This sheet entrains 
the mixture discharged from the first stage and carries it 




Fig. 188.— Steam-jet Type 
Vacuum Pump. 



CONDENSERS AND RELATED APPARATUS 289 



Steam Supply, First Stage 



Steam Strainer, 
First Stage 
h First Stage 
TSteam Nozzles 



First Stage 
Entrainment 
Space 

S 

'Air Suction 



Second Stage 
Steam Strainer' 




Collecting 
Space 



dscharge 



Fig. 189. — Cross-sectional View of the Radojet Vacuum Pump. 

into the diffusor shown. This diffusor consists of two cir- 
cular plates which are close together near the steam nozzle 
and separate gradually toward their circumferences. The 
mixture from the second diffusor is discharged into the 



290 



STEAM POWER 



collecting space shown and flows out at the discharge 
flange. 



Exhaust Steam Inlet 




Fig. 190. — One Method of Connecting Radojet to Surface Condenser. 



The radial flow from the second stage steam nozzle and 
through the second stage diffusor gives the device its trade 
name. 

In the arrangement illustrated in Fig. 189, the second 
stage steam has to entrain and compress not only the " air " 
coming from the condenser but also all of the steam used in 
the first stage. This requirement leads to the use of large 
quantities of second stage steam. When thermal efficiency is 
of sufficient significance, it is customary to separate the 
first and second stages by interposing an " inter cooler." 
This is merely a small surface condenser in which first stage 
steam and possibly vapors brought over from the condenser 
can be liquefied, leaving only saturated air to be handled 
by the second stage jet. 

When such an intercooler is installed, some or all of the 
condensate from the main condenser is generally used as 
circulating water, so that the heat given up by steam and 



CONDENSERS AND RELATED APPARATUS 291 

air is caught in Mi is condensate and returned to the 
boilers. 

119. Water Required by Contact Condensers. The 

weight of circulating water required varies with the type 
of condenser and with the conditions of operation, such as 
initial temperature of water, vacuum desired, etc. It can 
be determined approximately by calculation and the values 
thus found must then be increased by such factors as experi- 
ence has shown to be necessary. 

In contact condensers the water and the condensate 
are discharged as a mixture and therefore have the same 
average discharge temperature. 

Let ^1 = initial temperature of injection water in F.°; 

fe ~ temperature at which mixture is discharged in 

F.°; 
X = total heat above 32° F. of steam as exhausted; 
W = pounds of injection water per pound of exhaust 
steam. 

Assuming the exhaust steam to be dry saturated, each 
pound of steam in condensing to water at a temperature 
of fe degrees must give up an amount of heat equal to X 
minus the heat of the liquid at t°2 or roughly X— (fe — 32) 
B.t.u. This same quantity must be absorbed by the in- 
jection water, while its temperature rises from h to fa 
degrees. Each pound of water can then absorb approxi- 
mately fo — ti) B.t.u. and the pounds of injection water 
per pound of steam will be 

w _ \-k+S2 

i2 — h 

The value of fe would be that corresponding to the 
absolute pressure in the condenser if it were not for the 
air and similar gases which exert some pressure. It is 
generally 10 or more degrees F. below the temperature 
corresponding to the vacuum. Values of fe in the neigh- 



292 STEAM POWER 

borhood of 110° to 125° F. are customary with recipro- 
cating engines and values as low as 80° are used with high 
vacuums in connection with steam turbines. 

The weight of water used per pound of steam as given 
by Eq. (78) will vary between about 15 for very low initial 
and moderate discharge temperature to about 50 with 
average initial and moderate discharge temperature. Ex- 
perience shows that it is necessary to add 10 per cent or 
more to the values of W obtained from equation (78) to 
obtain the weight of water which will probably be used. 

ILLUSTRATIVE PROBLEM 

Find the quantity of water theoretically required per pound 
of steam condensed in a contact condenser in which a vacuum 
of 25.5 ins. of mercury is maintained when the barometer reads 
29.5 ins. of mercury. The initial temperature of the water is 
60° F. 

The absolute pressure in the condenser is 29.5—25.5=4.0 ins. 
of mercury and the steam tables give for this pressure, X = 11 15.0 
and to = 126. Substituting in Eq. (78) gives 

w 1115.0-126 +32 

^ = 7^ — ^ =15.5 approximately. 

126—60 

120. Weight of Water Required by Non-contact 
Condensers. In the case of non-contact condensers there 
is no definite relation between the discharge tempera- 
ture of the cooling water and that of the condensate. 
Experience shows that the discharge temperature of 
the circulating water is generally from 10 to 20 or more 
degrees lower than the temperature corresponding to the 
vacuum. 

The temperature of the condensate (hot-well tempera- 
ture) is often 15 or more degrees below that correspond- 
ing to the vacuum, but good design makes the hot-well 
temperature very closely approximate that corresponding 
to the vacuum. 






CONDENSERS AND RELATED APPARATUS 293 

Assuming 

^1 = initial temperature of injection water in F.°; 

i 2 = final temperature of injection water in F.°; 

t c = temperature at which condensate is discharged, i.e., 

hot- well temperature, in F.°; 
\= total heat above 32° F. of steam as exhausted, 

and 

W = pounds of injection water per pound of exhaust 
steam. 

The weight of water which must be circulated per pound 
of steam can be found as in the case of the contact con- 
denser. It is given by 

w = \-tc+Z2 (7Q) 

i2 — h 

Values in the neighborhood of 25 lbs. of water per pound 
of steam are common with low vacuums and 50 or more 
pounds are often used with vacuums over 28 ins. of 
mercury. 

ILLUSTRATIVE PROBLEM 

A surface condenser receives circulating water at a temper- 
ature of 65° F. and discharges it at a temperature of 80° F. It 
maintains a vacuum of 28.0 ins. with the barometer at 29.5, and 
the temperature of the condensate discharged to the hot well is 
equal to 85° F. Find the quantity of circulating water theoretically 
required. 

This vacuum corresponds to an absolute pressure of 
29.5—28.0 = 1.5 ins. of mercury. Assuming this all due to steam 
(neglecting presence of air) the value of X may be taken from the 
steam table as 1100.1 B.t.u. Substitution 'in Eq. (79) then gives 

w 1100.1-85+32 

" = en — az =69.9 approximately. 

oU — OO 

121. Relative Advantages of Contact and Surface Con- 
densers. The contact types are as a rule much cheaper 



294 STEAM POWER 

than the surface condensers, and they are less subject to 
serious depreciation, the tubes of surface condensers often 
corroding excessively in very short intervals of time. On 
the other hand, the injection of the cooling water into 
the condensing space in contact types results in the intro- 
duction of large quantities of dissolved gases, and much 
of this material is liberated under the reduced pressure, 
thus tending to increase the condenser pressure, that is, 
decrease the vacuum. Where pumps are used to carry 
away the mixture with contact condensers, these pumps 
have to handle a much larger quantity of water than the 
corresponding pump in a surface condenser, and the work 
of pumping this water out of the vacuum into the atmos- 
phere combined with the additional work required of 
the pump which handles the " air " may partly balance 
the advantage of lower first cost of the contact type. 

A surface condenser must always be installed where 
it is desirable to use the condensate as boiler feed, and 
it is generally used when very high vacuums (low absolute 
pressures) are to be maintained. The surface condenser 
is at a serious disadvantage, however, when required to 
handle the exhaust of reciprocating engines. The exhaust 
from such engines always contains large quantities of 
lubricating oil carried out of the cylinder, and unless this 
material is separated before the exhaust enters the con- 
denser it is deposited on the outer surfaces of the tubes 
and decreases the conductivity of those surfaces. Such 
oil can be eliminated to a great extent before the exhaust 
enters the condenser by means of oil separators, which are 
generally made up of a series of baffles upon which the 
steam impinges and upon which the oil is caught. 

122. Cooling Towers. The large quantity of circula- 
ting water required by condensing plants is often an item 
of great economic importance. When such plants are 
located near a river or near tide water, the circulating 
water can generally be procured for the cost of pumping. 



CONDENSERS AND RELATED APPARATUS 295 

When they are located in the middle of cities or in regions 
where water is scarce, the cost of water may be excessive 
or it may even be impossible to obtain a continuous supply 
equal to the demand of the condensers. 

In such cases the condensing water is often circulated 
continuously, being cooled after each passage through the 
condensers. This cooling is generally done by exposure 
of a large surface to the air. The resultant evaporation 
of some of the water with the absorption of its latent heat 
of vaporization cools the remainder so that it can be used 
again. This sort of cooling may be effected by running 
the water into a shallow pond of large area, or by spraying 
it into the air over a small pond or reservoir or by passing 
it through a cooling tower. 

Cooling towers are large wood or metal towers generally 
filled with some form of baffling devices. The hot water 
is introduced at the top and spread into thin sheets or 
divided up into drops as it descends. Air enters at the 
bottom and flows upward, cooling the water by contact 
and by the partial evaporation which results. The cir- 
culation of air may be natural, i.e., due to the difference 
of temperature between the. air inside and out, in which 
case a stack is fitted to the top of the tower; or the cir- 
culation may be forced by fans located in the base of the 
tower. In the latter case the apparatuses called a forced- 
draught cooling tower. 



CHAPTER XV 
COMBUSTION 

123. Definitions. Certain substances are known to 
chemists as compounds, because they can be separated by 
chemical processes into simpler substances. Thus water 
and many of the most familiar materials known to man 
are compounds which can be separated into two or more 
simpler materials. 

Those substances which cannot be further broken up by 
the processes used in separating compounds are called 
elements; they are regarded as elemental, as the stones 
of which the compounds of nature are built up. About 
eighty-three of these elements are now known, but many 
of them are comparatively rare. Pure metals are all 
elements; the oxygen and nitrogen which are mixed to form 
the greater part of the atmosphere are elements; carbon, 
which forms the greater part of most fuels, is an element. 

In many cases the combination of elements to form 
compounds is accompanied by the liberation of heat, and 
some of these combinations are used by the engineer for the 
purpose of obtaining heat in large quantities. When the 
elements which occur in fuels, such as coal, wood and 
petroleum, combine with oxygen, the process is spoken of 
as combustion. The quantity of heat liberated when a 
pound of any material combines with oxygen (burns) is 
called the heat value or calorific value of that material. 

Fuels contain a great number of elements, but only 
three of these ordinarily take part in combustion and are 
therefore spoken of as combustibles. They are carbon, 
hydrogen and sulphur. The sulphur content is generally 

296 



COMBUSTION 297 

very small, and the carbon and hydrogen are therefore the 
most important constituents. 

The combustion of each of these elements will be con- 
sidered in detail in the following sections, but before this 
can be done two other ideas must be developed. 

The smallest particle of an element which can be 
conceived of as entering into combination to form a com- 
pound is known as an atom of that element. It has been 
found that the atoms of each element have an invariable 
and characteristic mass. The lightest atom is that of 
hydrogen, and its weight is considered unity. The atom 
of carbon is twelve times as heavy as that of hydrogen and 
carbon is therefore said to have an atomic weight equal to 
twelve. Similarly the atomic weight of nitrogen is four- 
teen and that of oxygen is sixteen. 

The smallest particle which can be formed by the com- 
bination of atoms is known as a molecule. Like or unlike 
atoms may combine to form molecules. Thus two hydro- 
gen atoms combine to form a molecule of hydrogen, and 
hydrogen gas as it ordinarily exists may be pictured as 
made up of a collection of such molecules. Similarly, 
gaseous oxygen and gaseous nitrogen may be pictured as 
collections of molecules which are made up of two like 
atoms. 

When unlike atoms combine to form a molecule, they 
form a molecule of a compound. Obviously a molecule of 
any compound is the smallest particle of that compound 
which can exist. 

For convenience, the different elements are represented 
by abbreviations; thus oxygen is represented by 0, nitro- 
gen by N, hydrogen by H, carbon by C and sulphur by S. 
When these abbreviations are written in chemical equa- 
tions, such as will be given later, they stand for an atom 
of the substance. Hence in a chemical equation would 
mean one atom of oxygen. The symbol O2 is used to mean 
two atoms of oxygen in combination, hence, one molecule 



298 STEAM POWER 

of oxygen. The symbol 2O2 means two groups of two 
oxygen atoms in combination, hence two molecules of oxygen. 
The simplicity and elegance of this system will become 
apparent as the chemical equations which follow are de- 
veloped and explained. 

124. Combustion of Carbon. Carbon can unite with 
oxygen or burn to form two different compounds — carbon 
monoxide (CO) and carbon dioxide (CO2). The monoxide 
is formed by the combination of one atom of oxygen with one 
atom of carbon; the dioxide, by the combination of two 
atoms of oxygen with one of carbon. The dioxide, therefore, 
contains twice as much oxygen as does the monoxide. 

Carbon burned to carbon monoxide has not combined 
with the largest possible quantity of oxygen, and combus- 
tion is therefore said to be incomplete in such cases. When, 
however, carbon dioxide is formed, the carbon has combined 
with as much oxygen as possible and combustion is said to 
be complete. 

It will be shown later that much more heat is liberated 
when the dioxide is formed than when carbon burns to the 
monoxide. Hence, when liberation of heat is the object of 
combustion, the process should be so conducted as to result 
in the formation of the maximum quantity of dioxide and 
the minimum amount of monoxide. 

125. Combustion to CO. The combustion of carbon and 
oxygen to form the monoxide can be represented by the 
equation 

C+0 = CO, (80) 

or by the equation 

2C+0 2 = 2CO (81) 

The former is the simpler and will be considered first, but 
the latter is the more perfect and indicates more to the 
trained eye than does the simpler form. 

The simple equation states that one atom of carbon 
combined with one atom of oxygen to form one molecule 



COMBUSTION 299 

of carbon monoxide. It can, however, be so interpreted 
as to show much more than this. The carbon atom is twelve 
times as heavy as the hydrogen atom, while the oxygen atom 
is sixteen times as heavy as that of hydrogen. The equation 

C + = CO, 

therefore, shows that an atom, which is twelve times heavier 
than the hydrogen atom, unites with one which is sixteen 
times heavier than the hydrogen atom to form a molecule 
which is 28( = 12 + 16) times heavier than the hydrogen 
atom. 

In other words, the weights of carbon and oxygen 

12 3 1 

combining are in the ratio of t^ = t = 7T- If a sufficient 

lb 4 lg 

number of carbon atoms to weigh one pound be used, a 
quantity of oxygen weighing lj lbs. will be required to 
combine with them to form carbon monoxide. The re- 
sultant carbon monoxide will contain the pound of carbon 
and the 1J lbs. of oxygen and will therefore weigh 2f lbs: 

The same weight relations would hold irrespective of 
the weight of carbon used, and the simpler equation may 
therefore be put 

1 weight of C+l| weights of = 2^ weights of CO. (82) 

ILLUSTRATIVE PROBLEM 

To illustrate the use of this equation, assume that 9 lbs. of 
carbon are burned to carbon monoxide and that it is desired to 
find the weight of oxygen used, and the weight of the product. 
The weight of oxj^gen used must be 1^ times the weight of carbon, 
that is, 1^X9 = 12 lbs. The weight of the product must be 2\ 
times the weight of the carbon, that is 2f X9 =21 lbs.; or, it must 
be the weight of the carbon burned plus the weight of the oxygen 
used, that is, 9+12 =21 lbs. 

In general, the oxygen used for combustion is obtained 
from the atmosphere, which may be considered as a median- 



300 STEAM POWER 

ical mixture of oxygen and nitrogen in unvarying porportions. 

These proportions are roughly, 0.23 of oxygen to 0.77 of 

nitrogen by weight, or 0.21 of oxygen to 0.79 of nitrogen 

by volume, as shown in Table VIII. The weight of 

air which contains one pound of oxygen is therefore 

234-0 77 
' — = 4.35 lbs., and this weight of air contains 

4.35-1=3.35 lbs. of nitrogen. 

In the problem previously considered it was found 
that 12 lbs. of oxygen would be required to burn 9 lbs. 
of carbon to CO. The total weight of air required to 
obtain this oxygen will be 12X4.35 = 52.2 lbs. and it will 
contain 52.2 — 12 = 40.2 lbs. of nitrogen. 

By simple arithmetical calculations of the type just 
given all the weight relations involved in the combustion 
of C to CO can be determined. The volume of air required 
in any given case can be found by multiplying the weight 
of air by the specific volume as given in Table VIII. 
Thus, in the illustrative problem already considered, it 
was found that 52.2 lbs. of air would be required to burn 
9 lbs. of C to CO. The volume of this air at 62° F. would 
be 52.2X13.14 = 685.9 cu.ft. 

It is found that a quantity of heat equal to about 
4500 B.t.u. is liberated per pound of carbon burned to CO; 
that is the calorific value of C burned to CO is 4500 B.t.u. 

Returning now to Eq. (81), which was said to be more 
useful than the simpler form given as Eq. (80), it will be 
necessary to consider a rather simple law of gases. It 
has been shown experimentally that equal volumes of all 
gases contain the same number of molecules when at the same 
temperature and pressure. This statement is known as 
Avogadro's Law.. It has also been shown that the mole- 
cules of gaseous oxygen contain two atoms. 

The equation in question, 

2C+0 2 = 2CO 



COMBUSTION 



301 



can therefore be read, two atoms of carbon combine with 
one molecule of oxygen to form two molecules of carbon mon- 
oxide. But, if every molecule of O yields two molecules 
of CO it follows from Avagadro's law that the CO formed 
will occupy twice the volume of the oxygen used if measured 
at the same temperature and pressure. If the equation be 
imagined as containing a numeral 1 before the O2, it 
will be obvious that the coefficients of the terms represent- 
ing gas molcules give volume relations directly. This equa- 
tion therefore gives both volume and weight relations. 

TABLE VIII 
Properties of Air 

Considering it to consist only of nitrogen and oxygen. 





Relative Proportions. 


Ratio of N to 0. 


Ratio of Air to O. 




Exact. 


Approx. 


Exact. 


Approx. 


Exact. 


Approx. 


By weight . . 
By volume. . 


/ 0.766 N 
1 0.234 

/ 0.791 N 
10.209 


0.77N 
0.23 O 

0.79 N 
0.21O 


3.27 

3.78 


3.35 
3.76 


4.27 
4.76 


4.35 
4.76 





Spec. wt. at Atmos. Press. 
(Lbs. per Cu.ft.) 


Spec. Vol. at Atmos. Press. 
(Cu.ft. per Lb.) 




At 32° F. 


At 62° F. 


At 32° F. 


At 62° F. 




0.08072 


0.07609 


12.39 


13.14 



Weight of air containing one pound of oxygen is approximately 
4.35 lbs. 



126. Combustion to CO2. The combination of carbon 
and oxygen to form the dioxide is represented by the equa- 
tion 

C4-0 2 = C0 2? (83) 



302 STEAM POWER 

which shows that one atom of carbon (twelve times heavier 

than hydrogen) combines with two atoms of oxygen (each 

sixteen times heavier than hydrogen) to form a molecule 

of CO2, which is forty-four times heavier than an atom of 

hydrogen. Therefore the weight of carbon and oxygen 

12 3 1 
combining are as =^ = 7^2, so that 2| lbs. of oxygen 

are required to burn a pound of carbon to carbon dioxide. 
Writing this in the form of an equation, gives 

1 weight of C+2§ weights of = 3f weights of C0 2 . . (84) 

The weight of air required can readily be found by 
multiplying the required oxygen by the number 4.35, 
previously shown to be the number of pounds of air con- 
taining one pound of oxygen. Thus, the required air is 
2|X4. 35 = 11.57 pounds per pound of C burned to CO2. 
This number is commonly rounded out to 12 in engineering 
literature. 

The equation given shows volume relations directly. 
It is evident, therefore, that one molecule of O yields one 
molecule of CO2, and hence that the volume of the product 
is exactly equal to the volume of the oxygen used in forming 
it if measured at the same temperature and pressure. 
This is a very important relation, and is often made use 
of in engineering calculations. 

Experiment shows that when carbon burns to the 
dioxide about 14,600 B.t.u. are liberated per pound of 
carbon burned, that is, the calorific value of C burned to CO2 
in 14,600. 

127. Combustion of CO to CO2. Since carbon which 
has burned to carbon monoxide has not combined with the 
greatest possible quantity of oxygen, the monoxide can 
take up more oxygen by burning to the dioxide. This 
process is represented by the formula 

2CO+0 2 = 2C0 2 , (85) 



COMBUSTION 303 

which shows that two molecules of monoxide combine with 
one molecule of oxygen to form two molecules of the dioxide. 
The volume of CO2 formed is therefore equal to that of 
the CO burned. 

So far as the ultimate result is concerned, it makes no 
difference whether carbon is burned directly to CO2 or is 
first burned to CO and then the CO is burned to CO2. 
The total oxygen used per pound of carbon burned to CO2 
and the total heat liberated per pound of carbon burned 
to CO2 are the same in both cases. 

Thus, for the oxygen, one pound of C burned to CO2 
requires 2| lbs. of oxygen; but one pound of C burned 
to CO requires 1J lbs. of oxygen, and 1J lbs. more will be 
required when this CO is burned to CO2. The result 
is therefore the same. 

For heat liberated, one pound of C burned to CO2 
liberates about 14,600 B.t.u.; but one pound of C burned 
to CO liberates about 4500 B.t.u. and 10,100 B.t.u. are 
liberated when this CO is burned to CO2. Since the sum 
of 4500 and 10,100 is equal to 14,600 the result is again 
the same. 

Data on the combustion of C to CO and CO2 and the 
combustion of CO to CO2 are collected in convenient 
form in Table IX. 

128. Conditions Determining Formation of CO and CO2. 
Excluding certain complicated considerations which are 
not of great importance in steam-power engineering, it may 
be said that when carbon is being burned at a certain rate 
(pounds per unit "of time) the amount of oxygen brought 
into contact with ,the carbon determines whether the caibon 
burns to CO or to CO2. If enough or more than enough 
oxygen to burn the carbon to CO2 is brought into contact, 
that oxide will be formed. If there is not enough to burn 
all the carbon to the dioxide, both oxides are formed in cer- 
tain proportions, which can be calculated. 

Since combustion to CO yields only 4500 B.t.u. per 



304 



STEAM POWEK 



pound of C and combustion to CO2 yields 14,600 B.t.u. 
per pound of C, the importance of supplying sufficient 
oxygen to burn all carbon to the dioxide in cases where 
the liberation of the maximum quantity of heat is desirable 
is obvious. In actual practice the oxygen is furnished 
by supplying air and it is found necessary in most cases 
to supply more than the amount of air theoretically re- 
quired in order to insure burning all, or even nearly all, 
of the carbon to the dioxide. This comes from the great 
difficulty met in obtaining contact between the oxygen of the 
air and the carbon which is to be burned, that is, in bringing 
all the oxygen of the air into intimate contact with the 
fuel being burned in real apparatus. 



TABLE IX 

Combustion Data for Carbon 

(Per pound of carbon.) 



Product. 


Oxygen Required. 


Nitrogen Accompanying 
Oxygen. 


Pounds. 


Cu.ft. at 62° F. 
and 14.7 Lbs. 


Pounds. 


Cu.ft. at 62° F. 
and 14.7 Lbs. 


CO 


1.333 
2.667 
1.333 


16.0 
32.0 
16.0 


4.46 

8.92 
4.46 


60 1 


C0 2 fromC 

CO. from CO.... 


120.2 
60.1 





Air Required. 


Quantity of Product 
(N not included). 


Heat Liber- 
ated. 


Product. 


Pounds. 


Cu.ft. at 
62° F. and 
14.7 Lbs. 


Pounds. 


Cu.ft at 
62° F. and 
14.7 Lbs. 


CO 


5.79 
11.58 

5.79 


76.1 
152.2 

76.1 


2.33 
3.67 

3.67 


32.0 
32.0 

32. oj 


4,500 
14,600 
10,100 per lb. 
of C in CO 
4,300 per lb. 
of CO 


C0 2 from C 

C0 2 fromCO.... 



COMBUSTION 305 

The air in excess of that theoretically required to burn 
all the carbon completely is spoken of as excess air. In 
the form of an equation, this statement is equivalent to 

Air supplied — air theoretically required = excess air. (86) 

It is customary to express the quantity of excess air in 
terms of a numerical factor known as the excess coefficient. 
This coefficient is denned as the number by which the quantity 
of air theoretically required must be multiplied to give the 
quantity of air actually used. In the form of an equation 
this gives 

Excess coefficient X air theoretically required 

= air actually used. . (87) 

ILLUSTRATIVE PROBLEM 

Taking data from the illustrative problem previously considered, 
assume that 9 lbs. of carbon are burned in air to C0 2 . Each pound 
theoretically requires 11.57 lbs. of air, so that the theoretical 
air-supply for this case would be 9X11.57=104.13 lbs. If in a 
real case 150 lbs. of air are supplied, the excess coefficient is equal 
to 150-^ 104.13 =1.44. 

129. Flue Gases from Combustion of Carbon. The 

gases resulting from the combustion of fuels are known 
in engineering as the products of combustion or flue gases, 
because they are the gases passing through the flues or 
passages leading from furnaces in which fuel is burned and 
to the stacks which serve to carry off the gases. 

It has already been shown that the CO2 formed by the 
combustion of carbon has the same volume as the oxygen 
which is used in forming it. Therefore, if the air supplied 
in a given case just equaled that theoretically required 
for combustion to CO2 and if all of the oxygen were used, 
the CO2 formed would merely replace the oxygen in the 
air. The theoretical proportions of the flue gas would 
then be 0.21 of C0 2 and 0.79 of N by volume. 



306 



STEAM POWER 



If real flue gases obtained by burning carbon in air are 
found to contain less than 21 per cent of C0 2 , the combustion 
has evidently not yielded theoretically perfect flue gases. 
The trouble may be due to an excess or to a deficiency of air. 
If there is an excess of air there will be oxygen present in 
the flue gases; if there is a deficiency there will be CO 
present in the flue gases. An analysis of these gases for 
oxygen and for CO would therefore indicate the source of 
trouble and the remedy to be provided. 

The curve to the right of the central vertical line in 
Fig, 191 shows the theoretical decrease in volume per 



35 


\ 






















| 25 
> 

8 15 

> 

O 

£5 


\ 

\ 














































\ 

\ 


\ 






















'A 




















& 


. 


\ 








<% 
















\ 















50 40 30 



10 



Deficiency (in per cent) 



50 100 150 200 250 300 
Excess (in per cent) 

2 3 1 



Fig. 191. — Effect of Air Supply on Flue Gas Analysis. 



cent of CO2 in flue gases as the excess air increases. The 
single numbers 1, 2, 3 and 4 indicate the excess coeffi- 
cients corresponding to the various percentages of excess 
air. 

The curves to the left give the theoretical decrease in 
volume per cent of CO2 and the theoretical increase in 
volume per cent of CO as the air supplied is decreased below 
that theoretically required for complete combustion. 

130. Combustion of Hydrogen. Hydrogen combines 
with oxygen, or burns, to form water. The equation for 
this reaction is 

2H 2 +0 2 = 2H 2 0, ....*.. (88) 



COMBUSTION 307 

which indicates that two molecules of hydrogen combine 
with one molecule of oxygen to form two molecules of 
water. In terms of volumes, two volumes of hydrogen 
combine with one of oxygen to form two of gaseous water, 
that is, water in the form of highly superheated vapor. 
As the water is cooled down it will obviously approach 
and finally reach the liquid condition, with a rapid de- 
crease in volume quite different from that experienced 
by a gas under similar conditions, so that the volume rela- 
tions hold only at high temperatures. 

The weight relations can be calculated as in other 
cases, starting from the fact that four weights of hydrogen 
combine with thirty-two weights of oxygen to form 36 
weights of water. The weights of hydrogen and oxygen 
are therefore in the relation of -^ = |. 

The heat liberated when one pound of hydrogen burns 
to water is equal to about 62,000 B.t.u. This is the quantity 
of heat which could be obtained if one pound of hydrogen 
at, say, room temperature, and mixed with the requisite 
quantity of oxygen, were ignited and the resultant water 
were then cooled down to the initial temperature. During 
the cooling of the water it would partly or entirely condense 
and thus give up some or all cf its latent heat of vaporization. 
This heat would obviously be included in the calorific 
value just given. 

In many pieces of engineering apparatus in which 
hydrogen is burned the products of combustion are not 
cooled to such an extent that the water is condensed. The 
latent heat of vaporization would not be liberated under 
such conditions, but would remain bound up with the water 
vapor. When the water is not condensed the heat liberated 
is only about 52,000 B.t.u. per pound of hydrogen. This 
number is known as the lower calorific value of hydrogen, 
while 62,000 is known as the higher calorific value. 

Data on the combustion of hydrogen are given ip 
Table X. 



308 



STEAM POWER 



TABLE X 

Combustion Data for Hydrogen 

(Per pound of hydrogen) 





Oxygen Required. 


Nitrogen Accompanying Oxygen. 


Product. 


Pounds. 


Cu.ft at 62° F. 
and 14.7 Lbs. 


Pounds. 


Cu.ft. at 62° F 
and 14.7 Lbs. 


H 2 


8 


96 


26.8 


361 





Air Required. 


Quantity of Product (N 
not included). 


Heat 


Product. 


Pounds. 


Cu.ft. at 62° 

F. and 14.7 

Lbs. 


Pounds. 


Cu.ft. at 62° 

F. and 14.7 

Lbs. 


Liberated. 


H 2 


34.8 


457 


9 


Liquid 
0.144 


/ 62,000 
I 52,000 



131. Combustion of Hydrocarbons. Many of the fuels 
used by the engineer contain compounds of hydrogen 
and carbon which are called hydrocarbons. One of the best 
examples is methane (CH4), which forms the greater part 
of all the so-called natural gas. 

All of these hydrocarbons burn to CO2 and H2O if 
the supply of oxygen is great enough. If there is a deficiency 
of oxygen, combustion is incomplete and generally results 
in the formation of CO2, H2O, CO, C in the form of soot, 
and other products which need not be considered here. 

For complete combustion the requisite oxygen and 
air can be determined as in previous cases by means of 
chemical equations. Thus for methane the equation is 



CH 4 +202 = C02+2H 2 0, 



(89) 



which shows that sixteen (12+4) weights of methane 
combine with sixty-four (2X2X16) weights of oxygen to 
form forty-four (12+32) weights of carbon dioxide and 
thirty-six (4+32) weights of water. 



COMBUSTION 309 

The calorific value of hydrocarbons is generally assumed 
to be equal to the sum of the heat values of the carbon 
and hydrogen contained in one pound of the material. 
Thus, if C represent the fraction of a pound of carbon 
contained in one pound of the hydrocarbon and if H 
represent the fraction of a pound of hydrogen contained 
therein, the common assumption would make the higher 
calorific value of the hydrocarbon 

(C X 14,600) + (HX 62,000) B.t.u. . . (90) 

The results obtained in this way do not generally check 
well with the experimentally determined values, and it is 
best to use the latter when they are available. 

132. Combustion of Sulphur. Sulphur forms several 
different oxides, but when burned under engineering con- 
ditions it is generally assumed to form only the dioxide 
SO2. The chemical equation for such combustion is 

S+0 2 = S0 2 , (91) 

and since the atomic weight of sulphur is 32, this equation 
shows that equal weights of sulphur and oxygen combine 
to form the dioxide. 

The combustion of sulphur to SO2 liberates about 
4000 B.t.u. per pound of sulphur. 

133. Combustion of Mixtures. It is often necessary 
to obtain approximate calorific values of combustible 
materials which, without great error, can be considered 
as mixtures of combustible and non-combustible elements. 
If there is oxygen present in the mixture it is assumed 
to be combined with hydrogen in the form of water, so 
that the uncombined or available hydrogen per pound of 
material is given by the expression 

Available H = H-^, (92) 



310 STEAM POWER 

in which H and O respectively represent the fractions of 
a pound of hydrogen and oxygen in one pound of material. 
The calorific values of such a mixture containing car- 
bon, hydrogen and sulphur would then be given approxi- 
mately by the equation 



Higher B.t.u. = 14,6000+62,000 (h-^-) +4000S, 



(93) 



in which the letters stand respectively for the fractions 
of a pound of each of the elements present in one pound of 
the mixture. Similarly the lower calorific value would 
be (approximately) 

Lower B.t.u. - 14, 600C +52,000 /h-^) +4000S, (94) 
and the oxygen required will be 

Pounds of = 2fC+8(H-^)+S. . • (95) 



134. Composition of Flue Gases. It was shown in 
Section 129 that the flue gases resulting from combustion 
of carbon to carbon dioxide with the theoretical amount 
of air would consist of 21 per cent carbon dioxide and 79 
per cent nitrogen by volume. It was also shown that vari- 
ation from this composition in any real case can be inter- 
preted to show the cause of such variation. 

It is necessary to note that the figures given apply only 
to the case of pure carbon. If, for instance, hydrogen is 
burned with air, no carbon dioxide can result, since no 
carbon is present and the composition of the flue gases 
must therefore be quite different from what was indicated 
above. 

As a matter of fact, the flue gases resulting from the com- 
bustion of hydrogen with the theoretical air supply would, 
if cooled down to ordinary temperatures, consist of nitro- 
gen saturated with water vapor, All water vapor in excess 



COMBUSTION 311 

of that required to fill the space occupied by the nitrogen 
(at the existing temperature) would condense out as liquid 
water. 

All real fuels contain carbon and hydrogen and often 
sulphur as well. The theoretical composition of flue gases 
obtainable with real fuels is therefore quite different from 
that indicated for pure carbon. The theoretical air supply 
for such fuels contains just enough oxygen to burn the com- 
bustible constituents and the quantity of nitrogen which 
must accompany that oxygen. Irrespective of total quan- 
tities involved, the oxygen must represent about 21 per cent 
of the volume of the air and the nitrogen 79 per cent. 

After combustion is completed, some of this oxygen 
has been converted into carbon dioxide occupying the same 
volume as the oxygen from which it was made (when reduced 
to the same pressure and temperature). Obviously with 
real fuels the carbon dioxide must occupy less volume than 
all of the oxygen and therefore must form less than 21 per 
cent of the volume of the final products of combustion. 

The figures given below will serve to indicate the varia- 
tions which occur with real fuels. In each case a typical 
analysis has been assumed for the fuel named and it has 
beeen assumed that it is burned with the theoretical air 
supply. It should be noted that the percentage of carbon 
dioxide given holds only for the particular analysis of fuel 
assumed. 

Per Cent Carbon 
Fuel. Dioxide in 

Flue Gases. 

Bituminous Coal 18.4 

Wood 20.1 

Petroleum Oil 15 .4 

Natural Gas 11.7 

Blast Furnace Gas 25 . 1 

The high value for blast furnace gas is explained by the 
fact that this gas contains a large amount of carbon dioxide. 



312 STEAM POWER 

so that this quantity appears in the flue gas in addition 
to that formed by combustion of other constituents of the 
gas. 

135. Temperature of Combustion. If combustion of 
any material could be carried on inside an ideal vessel 
which did not absorb nor transmit heat, the heat liberated 
during the combustion could not escape from the space 
within the vessel. 

If the vessel contained initially only the combustible 
and the oxygen or air required to burn it, the products of 
combustion would be the only material contained within 
the vessel after the completion of combustion. Under 
such circumstances the heat would be used in raising the 
temperature of the products of combustion, and the process 
could be pictured as though all of the combustion occurred 
first, forming the products of combustion without change 
of temperature, and then the liberated heat raised the 
temperature of these products. 

Knowing the weight of each of these products and the 
quantity of heat required to raise the temperature of one 
pound of each of them one degree, the amount of heat 
required to raise all of them one degree could be found by 
multiplying the two known values. Thus, if carbon had 
been burned in oxygen to CO2 with the theoretical oxygen 
supply, the vessel would contain only carbon dioxide. 
To raise the temperature of one pound of carbon dioxide 
one degree requires an amount of heat equal to the specific 
heat of that gas. Therefore, if W represents the weight 
of CO2 formed and C represents its specific heat, the amount 
of heat required to raise the temperature of all of the CO2 
one degree would be W-C B.t.u. If Q B.t.u. were liberated 
by the combustion, the temperature rise in degrees would 
therefore be given by 

Temp, rise = t|L (96) 



COMBUSTION 313 

and if the initial temperature had been to degrees, the final 
temperature would be 

< = <o+^ (97) 



A final temperature figured in this way is called the theo- 
retical temperature of combustion. It can never be attained 
in practice because of heat lost to surroundings and because 
of other losses which need not be considered here. 

Theoretical temperatures of combustion are, moreover, 
nearly always calculated on the assumption that the specific 
heats of gases are constants, whereas they really increase 
with the temperature. It therefore follows that tempera- 
tures determined on the assumption of constant specific 
heat will be too high for this reason also. 

When gases are heated there are two distinctly different 
limiting possibilities; the volume occupied by the gases 
may remain constant or the pressure exerted by the gases 
may remain constant while the volume increases. In 
the case of constant volume all the heat added to the gases 
must be used for raising the temperature; the amount 
of heat required per pound per degree under these con- 
ditions is known as the specific heat at constant volume 
and is designated by C v . 

When, however, the volume is allowed to increase at 
such a rate as to keep the pressure constant the heat sup- 
plied must not only raise the temperature, but must also 
do whatever external work is done in displacing (pushing 
out of the way) surrounding mediums. The heat required 
per pound per degree under these conditions is known as 
the specific heat at constant pressure and is represented 
by C p . It is always greater than C v by the amount of heat 
required to do the external work accompanying a rise of 
temperature of one degree. 

Thus, in the case assumed above, had the vessel been 
so constructed that its internal volume did not change, 



314 STEAM POWEK 

the specific heat at constant volume would be used. On 
the other hand, had the vessel been fitted with a movable 
piston arranged to move outward at such a rate as to main- 
tain constant pressure within the vessel as the temperature 
rose, the specific heat at constant pressure would be 
used. 

In most cases the material burned is not pure carbon, 
but a fuel containing carbon, hydrogen and sulphur, and as 
air is generally used to furnish the oxygen, the products 
of combustion will contain not only the oxides of carbon, 
hydrogen and sulphur, but inert nitrogen as well. The 
temperature rise is determined in the same way, however, 
by dividing the heat liberated by the amount required to 
raise the temperature of the products one degree. Thus 
if Wij W2, Ws . . . W n stand for the weights of the various 
products and C\, C2, C3 . . . C n for their respective specific 
heats, the theoretical temperature rise is given by 

Temp.rise = TFiCi+ ^ C2 + ^ C3+ _ _ ^^ (98) 
and the theoretical temperature of combustion is given by 

t==t ° + W 1 C 1 + W 2 C2 + WsCs . . . W n Cn' ^ 

if to stands for the initial temperature. 

PROBLEMS 

1. Assume 10 lbs. of C burned to CO. Determine the quan- 
tity of oxygen required, the quantity of air required, the quantity 
of nitrogen in this air, and the quantity of heat liberated. 

2. What will be the volume of the CO formed as above if 
measured at 62° F. and 14.7 lbs. pressure? 

3. Assume 15 lbs. of C burned to C0 2 . Determine the quan- 
tities of oxygen and air required, the quantity of nitrogen con- 
tained in this air, and the quantity of heat liberated. 

4. What will be the volume of the C0 2 formed from 15 lbs. 
of carbon if measured at 62° F. and 14,7 lbs.? 



COMBUSTION 315 

5. What will be the volume of the flue gases formed by the 
combustion of 11 lbs. of carbon to C0 2 with the theoretical air 
supply? 

6. The quantity of CO obtained by the combustion of 8 lbs. 
of carbon is burned to C0 2 with the theoretical amount of oxygen. 
Determine the quantities of ox}^gen and air required, the amount 
of nitrogen contained in this air, and the quantity of heat liberated. 

7. Assume 5 lbs. of C burned in air to C0 2 with an excess 
coefficient of 1.5. Determine the quantities of oxygen and air 
supplied, the heat liberated and the composition of the flue gases. 

8. The composition of flue gases resulting from the combus- 
tion of carbon in air is found to be 21% of C0 2 and 79% of N 
by volume. What is the value of the excess coefficient? 

9. An analysis of flue gases resulting from the combustion 
of carbon in air shows 12% of C0 2 by volume and no CO. The 
gases are not analyzed for or N. What can you say with regard 
to the air supply? 

10. Three pounds of hydrogen burn with theoretical oxj'gen 
supply. Determine the weight of oxygen and air used, the weight 
of the resultant water and the weight of the flue gas. 

11. Determine the heat liberated in the preceding problem if 
the water vapor is condensed and if it is not condensed. 

12. How much hydrogen would have to be burned to obtain 
20 lbs. of water? 

13. The chemical formula of methane is CH 4 . If one pound 
of methane is burned with theoretical air supply, what weight 
of air will be used, and what will be the weight of the flue gases? 

14. What would be the percentage composition of the flue 
gases of the preceding problem on a weight basis? 

15. The chemical formula of ethane is C 2 H 6 . Determine the 
calorific value of this material by means of the formula given 
in the text. 

16. A certain material is found to have the following analysis 
on a weight basis: C, 85%; H, 12%; 0,1%; S, 2%. Determine 
the calorific value of this material by means of the formula given 
in the text, assuming that all the oxygen present is in combination 
with hydrogen. 

17. Determine the amount of oxygen required to completely 
burn 3 lbs. of the material described in the preceding problem. 

18. One pound of carbon is burned to C0 2 in pure oxygen 
hi a vessel so arranged as to maintain constant internal pressure. 
The specific heat of C0 2 at constant pressure and at ordinary 
temperatures is about 0.21. Calculate the theoretical temperature 



316 STEAM POWER 

rise and the temperature of combustion, using this value of the 
specific heat and assuming an initial temperature of 60° F. 

19. Make the same calculations as called for in the preceding 
problem, but using the value 0.27 for the specific heat of C0 2 . 
This is more nearly the average value of the specific heat over 
the range of temperature existing in such a case. 

20. The hydrocarbon ethylene is represented by the chemical 
formula C 2 H 4 . Assume that one pound of this material is burned 
in air within a vessel arranged to maintain the products at con- 
stant pressure and that the excess coefficient is 1.5. Determine 
the theoretical temperature of combustion if the initial temperature 
>s 60° F., the mean specific heat of C0 2 is 0.27, that of H 2 is 

,61, that of N is 0.27, and that of is 0.24. 



CHAPTER XVI 

FUELS 

136. Commercial Fuels. In engineering practice any- 
thing which is combustible and which can be procured in 
large quantities at a reasonable cost is called a fuel. The 
principal commercial fuels are: 

f(l) Coal. 
a. Solid 1 (2) Wood and wood wastes. 
1(3) Vegetable wastes. 

{(1) Crude petroleum or natural oil. 
(2) Various products made from petroleum. 
(3) Methyl and ethyl alcohol. 

(1) Natural gas. 

(2) Artificial or manufactured gases. 



c. Gaseous 



Coal is by far the most extensively used fuel because of 
its abundance and relative cheapness in most localities. 
However, in oil-producing regions the crude oil and some 
of the products made from it are more often the commonly 
used fuel, particularly if good coal is not mined in the 
immediate vicinity. 

Wood is, in general, too valuable to be used exclusively 
as a fuel excepting on the frontiers where wooded terri- 
tory is being opened up and where coal cannot be pro- 
cured excepting at prohibitively high cost. Wood wastes, 
on the other hand, are very often used for fuel in the indus- 
tries producing them. 

Vegetable wastes, like wood wastes, are essentially of 
local value, being practically entirely consumed by the 
industries producing them. 

317 



318 STEAM POWER 

Natural gas is in many respects an ideal fuel, and is 
extensively used for power production in some localities. 
The diminution in the quantity available, the consequent 
rise in the price, the great economy achieved in burning 
this gas in gas engines and the increased use of the gas for 
domestic purposes are, however, gradually eliminating this 
fuel from the steam-power field. 

Artificial gases have never been extensively used for the 
generation of steam, as it is generally cheaper to burn the 
materials from which the gases are made, rather than to 
convert them into gas and then to burn the gas under 
boilers. This condition may change in the future when 
better markets have been opened up for some of the by- 
products which can be obtained from artificial gas plants. 

137. Coal. The word coal is used as the name of a 
great group of natural fuels which consist of more or 
less metamorphosed vegetable remains. At one end of the 
group is the material known as peat, which is only slightly 
changed from the original vegetable substance; at the 
other end is the graphitic anthracite which has undergone 
such radical metamorphosis that practically all of the 
original vegetable material excepting carbon and ash has 
been eliminated. 

A common, rough classification of the coals in the order 
of age, or of completeness of carbonization is, 

1. Peat or turf. 

2. Lignite (brown or black). 

3. Sab-bit aminous coal. 

4. Bituminous coal. 

5. Semi-bituminous coal. 

6. Semi-anthracite. 

7. Anthracite. 

8. Graphitic anthracite. 

The divisions are not at all exact, as they depend partly upon 
chemical composition and partly upon physical properties. 



FUELS 



319 



Another classification of a more exact variety is that 
given in Table XI and partly illustrated in Fig. 192, which 
gives what is known as Mahler's curve. It is for United 
States coals only. The terms used in this classification 
are explained in subsequent paragraphs. 




60 70 80 90 

$Fixed Carbon in the Combustible 



100 



Fig. 192. — Mahler's Curve for United States Coals. 

TABLE XI 

Classification of Coals 



Division. 



Graphitic 

Anthracite 

Semi-anthracite .... 
Semi-bituminous. . . 
Eastern bituminous. 
Western bituminous 
Lignite 



Per cent of 

Fixed Carbon 

in Combustible 



100 to 97 
97 to 92.5 
92.5 to 87.5 
87.5 to 75 
75 to 60 
65 to 50 
under 50 



Per cent of 
Volatile Matter 
in Combustible. 



to 3 

3 to 7.5 

7.5 to 12.5 

12.5 to 25 

25 to 40 

35 to 50 

over 50 



Calorific Value, 

B.t.u. per Pound of 

Combustible. 



14,600 to 
14,900 to 
15,300 to 
15,600 to 
15,800 to 
15,200 to 
13,700 to 



14,900 
15,300 
15,600 
15,900 
14,800 
13,700 
11,000 



The graphitic anthracite occurs in very small quantities 
and mostly in Rhode Island, With a few minor exceptions 
the anthracites occur only in Eastern Pennsylvania and the 



320 STEAM POWER 

semi-anthracites are almost entirely confined to the western 
edge of this field. 

The semi-bituminous coals are found on parts of the 
eastern border of what is known as the Appalachian coal 
field, extending from central Pennsylvania through the 
intermediate States to the northern part of Alabama. The 
greater part of this enormous bed consists of eastern bitu- 
minous coal. Western bituminous coals are found in large 
beds in the central part of the United States, principally in 
the States of Illinois, Indiana and Kentucky on the east of 
the Mississippi River, and in Iowa, Kansas and Texas to 
the west of that river. 

Lignite is found in small quantities in nearly all of the 
western half of the United States and in large beds in the 
Dakotas, Texas, Arkansas, Louisiana, Mississippi and 
Alabama. 

Peat is distributed in small beds throughout practically 
all of the United States and is continually forming in many 
marshes and on low-lying lands. 

138. Coal Analyses. Two different coal analyses are 
in use, the simpler being known as the proximate analysis 
and the more exhaustive being called the ultimate analysis. 
Both are made and reported on a weight basis. 

The proximate analysis assumes coal to contain four 
different and separable things, which are called fixed carbon, 
volatile hydrocarbon or volatile matter or volatile, moisture 
and ash. 

Moisture is determined by maintaining a small quantity 
of finely ground coal at a temperature of about 220° F. 
for one hour. The material lost during this time is 
assumed to be moisture only and is reported as such. 
Coal from which the moisture has been driven in this way 
is called dry coal. 

Volatile matter is determined by heating a sample 
from which the moisture has been driven, or a fresh sample. 
The coal is maintained at a red to white heat with exclu- 



FUELS 321 

sion of air until there is no further loss of weight. In 
the case of a previously dried sample the loss under these 
conditions is called volatile hydrocarbon. If the sample 
was not previously dried a separate moisture determina- 
tion is made on a similar sample and the weight of volatile 
is found by difference. 

Fixed carbon is found by combustion of a sample from 
which the moisture and volatile have been driven, the 
loss under these conditions being assumed to be entirely 
due to the combustion of carbon. 

Ash is the name given to the incombustible material 
left behind after determining the fixed carbon. 

The volatile hydrocarbons and the fixed carbon as 
determined in the proximate analysis are assumed to be 
the only combustible parts of the coal and their sum is 
called the combustible. 

Proximate analyses are reported in three different 
ways: On coal as received, on dry coal, and on combustible. 

Since the water content of a sample of coal received 
at any plant is largely a matter of the weather conditions 
during shipment, the best idea of the character of a coal 
can be obtained by excluding the consideration of its 
moisture content. It is generally best, therefore, to convert 
analyses to a dry coal basis, that is, recalculate the per- 
centages of volatile, fixed carbon and ash on the assumption 
that the analysis was made on the weight of coal which would 
result from drying the sample that was actually used. Ex- 
cessive moisture is, however, undesirable for steam-raising 
purposes, and the amount of moisture should therefore be 
determined in every case. 

Ash is also more or less a matter of accident in that the 
amount contained is largely determined by the care used 
in mining and subsequent cleaning of the coal. While 
it has a very appreciable effect upon the character of the 
material as a fuel it really has little connection with the 
combustible part of the fuel. For purposes of classifica- 



322 STEAM POWER 

tion, therefore, the ash should also be eliminated and the 
analysis given on the basis of combustible. 

Sulphur is sometimes reported with a proximate analy- 
sis. In making such an analysis the greater part of the 
sulphur is really driven off with, and regarded as, part of 
the volatile, so that when the sulphur content is desired it 
must be determined by a separate analysis. 

The ultimate analysis attempts to separate the dry 
combustible into the various elements of which it is com- 
posed. The percentages of carbon, hydrogen, oxygen, 
nitrogen and sulphur are determined as well as the per- 
centage of ash in dry coal. Such analyses show the carbon 
contents of coal to vary from about 98 per cent in the 
graphitic anthracite through about 97 per cent in the 
semi-anthracite, 87 per cent in semi-bituminous, 80 per cent 
in bituminous and 74 per cent in lignites to as low as 61 
per cent in peats. The corresponding figures for hydrogen 
run from about 1 per cent through a range in the neigh- 
borhood of 5 per cent for semi-bituminous to about 6 per 
cent in the case of peat. 

Oxygen varies from about 2 per cent or less in the 
case of good anthracite to as high as 33 per cent for peat; 
nitrogen generally forms about 1 per cent of the dry fuel 
and sulphur from 1 to 3 per cent. 

139. Calorific Value of Coals. The calorific value of 
coals on a basis of combustible has been shown to vary 
approximately according to a smooth curve, but the local 
variations are so great that no generally applicable formula 
for calorific value has yet been proposed. The formula 
commonly used is based upon the ultimate analysis and 
is similar to that suggested as approximately applicable 
in the case of mixtures of combustibles. It is known as 
Dulong's formula, and is 

[62,0001 / n \ 
B.t.u. perlb. = 14,600C + or ( H-^-J +4000S, (100) 

[52,000 J V */ 



FUELS 323 

in which the letters refer to the weight of the various ele- 
ments contained in one pound of dry coal. 

When an accurate knowledge of the calorific value of a 
fuel is desired it should be obtained by means of a fuel 
calorimeter. There are many varieties of this instrument, 
but practically all operate on the same general principle. 
A known weight of fuel is completely burned within a vessel 
and the heat liberated is absorbed by water or similar 
liquid. From measurements of liquid temperatures the 
heat absorbed by the liquid can be determined, and this 
with some additions for losses of various kinds must be the 
heat liberated by the fuel. For details see Chapter XX. 

140. Purchase of Coal on Analysis. Until quite recently 
it was customary to buy coal from the lowest bidder pro- 
vided the material supplied could be made to give satis- 
factory results in the plant. Obviously the purchaser knew 
nothing regarding his purchase, and often bought quantities 
of ash and moisture at the price of combustible. Now, 
however, the larger power plants and many of the smaller 
are buying on the basis of analyses and calorific values as 
determined in calorimeters. 

A certain desirable standard analysis is set and cer- 
tain variations are allowed from it. Wide variations are 
penalized by deducting so many cents per ton for each 
variation of a certain degree, and, finally, outside limits 
are set for moisture and ash beyond which the fuel need not 
be accepted. In some cases limits are also set for sulphur. 

This is the logical method of purchasing coal in large 
quantities, and is sure to come into very general use as its 
advantages become known. 

141. Petroleum. This material is obtained from drilled 
wells and has been found in many widely separated sections 
of the country. The oil wells of Pennsylvania and neigh- 
boring States, of Oklahoma, Texas and California have been 
the most productive and are hence the most widely known. 

Natural petroleum, as it occurs in the United States, is 



324 



STEAM POWER 



generally a dark, rather thick, oily liquid with a char- 
acteristic odor. It varies widely in composition so far as 
the compounds contained are concerned, but the variations 
in ultimate composition, specific gravity and calorific value 
are comparatively small. 

The ultimate analysis of crude oil generally shows about 
83 to 85 per cent of carbon, 13 to 15 per cent of hydrogen 
and small quantities of oxygen, nitrogen and sulphur. 

The specific gravity generally lies between 0.80 and 0.90 
and in most cases is nearer the upper figure. It is common 
practice to express the gravity in terms of the Beaume 
scale, an arbitrary scale developed for an instrument known 
as the Beaume hydrometer. This device is arranged to 
float in liquids and measures the gravity by the distance to 
which it sinks. Various corresponding values of the Beaume 
scale and specific gravity are given in Table XII for the 
region most used in connection with petroleum. 

TABLE XII 
Corresponding Beaume Readings and Specific Gravities 



Beaum6 Reading. 


Specific Gravity. 


Beaume' Reading. 


Specific Gravity. 


20 


0.9333 


34 


0.8536 


22 


0.9210 


36 


0.8433 


24 


0.9090 


38 


0.8333 


26 


0.8974 


40 


0.8235 


28 


0.8860 


42 


0.8139 


30 


0.8750 


44 


0.8045 


32 


0.8641 


46 


0.7954 



The higher calorific value varies between 19,000 and 
20,000 B.t.u. per pound and the lower value is generally 
1000 to 1500 B.t.u. lower. 

Crude oil is sometimes used for fuel, but this is unde- 
sirable, for two reasons. First, the crude oil contains 
many highly volatile constituents which can be distilled 



FUELS 325 

off and which have a high market value in the forms of 
gasoline and allied distillates. Second, the presence of 
these highly volatile constituents in the oil makes it more 
dangerous, as combustible vapors are given off in large quan- 
tities at low temperatures and the mixtures formed with 
the oxygen of the air are often highly explosive. 

As a consequence, the material generally sold as fuel 
oil is a residuum left after distilling off the more volatile 
constituents of the crude oil. It has practically the same 
properties as the crude, excepting that dangerous vapors 
are not given off at so low a temperature. 

PROBLEMS 

1. A sample of coal gives the following proximate analysis: 
moisture, 5%; volatile, 4.25%; fixed carbon, 80.75%; and 
ash, 10%. Determine the percentage of combustible and the 
percentages of fixed carbon and of volatile in the combustible. 

2. What variety of coal is indicated by the values obtained 
in Prob. 1? 

3. The following results were obtained in making a proximate 
analysis of a sample of coal; moisture, 7%; fixed carbon, 56.7%; 
volatile, 24.3%; ash, 12%. Determine the percentage of com- 
bustible and the percentages of fixed carbon and of volatile in the 
combustible. What variety of coal is indicated by these values? 

4. The ultimate analysis of a sample of dry coal gave the 
following results: carbon, 79.12%; hydrogen, 4.14%; oxygen, 
1.84%; sulphur, 0.92% ; nitrogen, 0.74%; ash, 13.24%. Recal- 
culate these values for an ash-free coal. 

5. Determine by means of Dulong's formula the upper and 
lower calorific values of the coal described in Prob. 4. 

6. The ultimate analysis of a sample of crude petroleum from 
which all water was removed gave the following results: carbon, 
85%; hydrogen, 13%; sulphur, 1.0%; oxygen, 0.25%; nitrogen, 
0.12%; ash (sand and similar material), 0.63%. Determine the 
upper and lower calorific values by means of Dulong's formula. 



CHAPTER XVII 
STEAM BOILERS 

142. Definitions and Classification. The term boiler 
is generally applied to the combination of a furnace in which 
fuel may be burned continuously and a closed vessel in which 
steam is generated from water by the heat liberated within 
the furnace. 

Boilers are classified in many different ways, the more 
important being given in the following schedule: 

Classification of Boilers 



(1) According to form 



"(a) Plain cylindrical, 

(6) Flue, 

(c) Tubular, 

(d) Sectional, etc. 



(2) According to location of J (a) Externally fired, and 
furnace [ (b) Internally fired. 



(3) According to use 



(4) According to direction of 
principal axis 



(a) Stationary, 

(6) Portable (as on trucks, 
or rollers), 

(c) Locomotive, 

(d) Marine. 

(a) Horizontal, 

(6) Inclined, 

(c) Vertical. 



(5) According to relative posi- f 

tions of water and hot I 

gases I 

326 



(a) Water tube, 

(b) Fire tube. 



STEAM BOILERS 



327 



Examples of boilers of the different types mentioned are 
given in subsequent paragraphs, 

143. Functions of Parts. It has been shown that there 
are two essentially different parts in the apparatus commonly 
known as a steam boiler, the furnace and the boiling vessel. 
A simple form of boiler known as a horizontal, return tubu- 
lar boiler, or an H.R.T. boiler, is shown in Figs. 193 and 
194 with the two essential parts and their components 

Pressure Regulator 



M 



•. • .'— j 



Steam 
Dome 



cr ": ~:~- 




Fig. 193.— Sectional Elevation of H.R.T. Boiler and Furnace. 



indicated. The furnace consists essentially of the combina- 
tion of grates, bridge wall, fire and ash doors, the ash pit 
and the space above the grates. It is the function of the 
furnace to so burn the fuel that the maximum amount 
of heat will be made available for absorption by the water 
within the boiling vessel. 

It is the function of the boiling vessel to transmit to 
the water within it the greatest possible quantity of the 
heat thus made available and to resist successfully the 



328 



STEAM POWER 




zm A 



tendency to rupture under the action of the high internal 
pressure, that is, the pressure of the steam. 

In the type of boiler shown the fuel is " fired " by 
hand, that is, it is 
spread on the grate by 
being thrown from a 
scoop shovel through 
the opened fire door. 
Air enters through both 
doors in regulated pro- 
portions and in such 
quantities as best to 
approximate complete 
combustion. 

The hot gases re- 
sulting from the com- 
bustion pass over the 
bridge wall, along the Fig. 194. 

lower part of the boiler Section through Furnace of H.R.T. Boiler, 
shell and then through 

the fire tubes, or flues, toward the front of the boiler as 
shown by arrows in the figure. From the front end of the 
tubes the products of combustion pass up through the 
smoke box to " breechings " or " flues," which carry them 
to the stack. 

Heat is received by the water within the vessel in two 
different ways: 

(1) The hot fuel bed on the grate radiates energy in the 
same way that the sun or any other glowing body radiates 
energy. Some of this energy traverses the space between 
fuel bed and boiler shell and ultimately passes through that 
shell to the water within. The rest of the radiated energy 
passes into the walls surrounding the furnace and heats 
them and the surrounding atmosphere. 

(2) The hot gases of combustion pass over the heating 
surface of the boiler, as shown, and transmit part of their 



STEAM BOILERS 329 

heat to the water on the other side of those surfaces. The 
rest of the heat which they carry is either lost to the surround- 
ing walls or is carried up the stack by the gases which 
leave the boiler at a comparatively high temperature. 
This temperature ordinarily ranges from about 500° to 700° 
F. and in extreme cases goes even higher. 

144. Furnaces and Combustion. In most forms of 
boiler the water within the boiler has practically the same 
temperature as the steam being generated, and this is 
generally from 320° to 400° F. Obviously the products 
of combustion cannot be cooled by the water to a tem- 
perature below that of the water, so that the gases leaving 
the boiler in an ideal case would have a comparatively 
high temperature. Practically, it is found undesirable to 
attempt to reduce the temperature of the gases to a value 
even approximating that of the water and, as indicated 
above, they are discharged at a temperature several hundred 
degrees higher. In order that the maximum amount of 
heat may be made available for the boiling vessel the prod- 
ucts of combustion must therefore leave the furnace with 
the highest possible temperature, and the ideal furnace 
would completely burn the cheapest fuel available in such 
a way as to give this highest possible temperature and not 
to generate smoke. 

Real furnaces fall far short of this ideal performance, 
for numerous reasons. The more important of these are 
given in the following paragraphs: 

(a) Incomplete Combustion of Carbon. In a real furnace 
the combustion of the carbon of the fuel may be incomplete 
in two senses; first, some of the carbon may remain entirely 
unoxidized and pass off with the ash, and second, some of 
the carbon may be burned to CO instead of to CO2. 

Imperfect combustion of the first kind can result from 
fuel falling through the openings in the grate before it has 
been ignited or when only partly burned, or it can result 
from failure to get air to some of the carbon in sufficient 
quantities to burn it completely before all of the surround- 



330 STEAM POWER 

ing fuel has been converted into ash and the locality 
cooled down to such an extent as to allow the unburned 
carbon in its midst to cool below the temperature of 
ignition. 

Imperfect combustion of the second kind, resulting in 
the formation of CO, generally results from a lack of suffi- 
cient air or from imperfect mixing of air and fuel or of air, 
fuel and products of combustion. It can also be caused by 
chilling of air or gases rising from the fuel bed before com- 
bustion has been completed. In any coal-burning furnace 
there is a tendency toward the formation of CO within the 
bed of fuel, and this tendency is greater the greater the 
depth of the bed. Such CO burns to CO2 above the fuel 
bed when conditions are propitious, the necessary air either 
passing through the fuel or being admitted at a point above 
the fuel. It is obvious that if combustion of such gas is 
to occur in this way, the combustible gas and the air must 
be brought into intimate contact while they are still at a 
sufficiently high temperature. 

The percentage of CO2 in the flue gases is commonly 
taken as an indication of the character of combustion. 
With pure carbon and theoretical air supply, this should 
be 21 per cent, and with real coal something between 18 
and 20 per cent. Practically, it is so difficult to bring the 
combustible material and the oxygen of the air into inti- 
mate contact that a large excess of air is always used. The 
excess coefficient in practice varies from about 1.1 to over 
2 with averages of about 1.5 under ordinary good conditions. 
The value 1.5 corresponds roughly to about 13 per cent 
CO2 in the flue gases and an excess coefficient of 2 cor- 
responds roughly to 9 to 10 per cent CO2. Even with such 
excess coefficients as those indicated as averages, it is not 
at all uncommon to find a small quantity of CO in the flue 
gases. The fireman's task with respect to the combustion 
of carbon thus reduces itself to the use of such a quantity 
of air in such a way that the minimum loss results from 
excess air arid Unburned CO. 



STEAM BOILERS 331 

(b) Incomplete Combustion of Hydrocarbons. The hydro- 
carbons which appear as volatile matter in the proximate 
analysis are practically all distilled from the fuel, as it is 
heated in the furnace before ignition in the same way as 
when making a proximate analysis. If they are to be 
completely burned they must be mixed with the requisite 
quantity of air after distillation and both the vapors and the 
air must be maintained at a sufficiently high temperature 
until combustion is complete. The air for the combustion 
of distilled volatile may all filter through the fuel bed or 
some of it may be admitted at a point above the level of 
the fuel bed. 

If the flame formed by burning hydrocarbons is allowed 
to come in contact with cold surfaces, as, for instance, 
the heating surfaces of the boiler, the gases are cooled 
below the temperature of ignition and combustion ceases. 
This results in the deposit of soot (unburned carbon) upon 
the heating surfaces of the boiler and in the carrying of 
soot and unburned hydrocarbons up the stack. The soot 
and some of these hydrocarbons form the unsightly smoke 
so familiarly associated with some stacks. 

Or, if the air supply is at a sufficiently high temperature, 
but is insufficient in quantity, the hydrocarbons are in- 
completely burned and smoke results. 

The formation of smoke can be conveniently studied 
by means of the ordinary kerosene lamp. Such a lamp 
operates by burning hydrocarbons of the same general 
character as those distilled from solid fuels. The hydro- 
carbons are drawn up by the wick in the form of liquids, 
are vaporized by heat near the top of the wick and then 
combine with oxygen from the atmosphere to give the 
luminous kerosene flame. 

If the flow of kerosene and the air supply are properly 
adjusted and if the temperature is high enough, the com- 
bustion results in the formation of invisible and practically 
odorless gases. If, however, the air supply be decreased 



332 STEAM POWER 

or be greatly cooled, a very smoky and very odorous combus- 
tion ensues. The same result could be obtained by the use 
of too great a quantity of air, a condition often attained 
when the supply of kerosene in the bowl of the lamp is almost 
exhausted. 

The effect of a cold surface is easily seen by inserting 
a cold metallic or porcelain surface into the tip of the flame 
and then withdrawing it. It will be found covered with 
soot. 

(c) Advantages and Disadvantages of Excess Air. It 
has been shown that excess air is practically necessary in 
the real furnace in order to insure against a deficiency at 
any point, and it is thus advantageous in that it makes the 
combustion more nearly complete than would otherwise 
be the case. On the other hand, excess air represents just 
so much excess material to be heated at the expense of heat 
liberated by combustion and hence decreases the maximum 
temperature attained. A sufficiently great supply of excess 
air could so reduce the temperature that even if combus- 
tion were complete very little heat would be made available 
for absorption by the boiling vessel, because the temperature 
attained by the products of combustion would be too low. 

Excess air in large quantities may also result in cooling 
unburned gases before combustion to such an extent as to 
make the completion of combustion impossible. 

145. Hand Firing. The commonest type of furnace is 
that shown in Figs. 193 and 194, and the commonest method 
of hand firing consists in spreading a layer of fuel as evenly 
as possible over the entire surface of the fuel bed as often 
as required to replace the fuel burned away. At such inter- 
vals as experience shows to be necessary the fire is cleaned, 
that is, the ashes are worked out from under the fuel by 
means of slice bars, so that practically nothing but live 
fuel resting on a thin layer of ash remains behind. 

This method is open to many serious objections; the 
more important are ; 



STEAM BOILERS 333 

1. There is a gradual increase in thickness of fuel bed 
from the time of one cleaning until the time of the next. 
This gives a constantly changing set of requirements for 
the proper proportions of air entering below and above 
the fuel bed and a constantly changing resistance to flow 
of air through the bed, so that great skill is necessary if the 
best conditions are to be maintained throughout. 

2. There is always a tendency for a fuel bed to burn 
faster at some points than at others, due to the accidental 
distribution of fuel, ash and air. Where " holes " are 
formed in this way large quantities of comparatively cold 
air can pass through with the consequences already enumer- 
ated. It takes considerable skill and watchfulness on the 
part of the fireman to prevent the formation and continued 
existence of such holes. 

3. The firing door must be opened wide every time 
that fuel is to be fired, that is, at intervals varying from two 
or three minutes to fifteen or more, depending on load, 
character of fuel, etc. While the door is open large quanti- 
ties of cold air readily flow into the furnace and cool down 
all parts of it, and a proportionately smaller amount will 
ordinarily pass through the fuel bed. The result of this on 
the flue gases and operation of the boilers has already been 
considered, but there is another result of equal or greater 
importance. As a consequence of this action the volatile 
hydrocarbons distilled off from the freshly fired fuel, which 
are themselves at a comparatively low temperature, are 
surrounded on all sides by cooled walls and come in contact 
with cold air only. The chances of their burning completely 
are very slight, and a great part of these volatilized materials 
passes off unburned as invisible gas and as smoke. Ob- 
viously the greater the volatile content the greater the dif- 
ficulty, so that anthracite causes least trouble in this way, 
while most bituminous coals give heavy black smoke when 
burned under these conditions. 

The cooling down of the interior of the furnace during 



334 STEAM POWER 

firing is accompanied by the covering of the fuel bed with 
cold fuel, so that, for the time being, very little radiant 
heat enters the boiling vessel, and the gases which come in 
contact with its surface are comparatively cool. The 
maintenance of a constant steam pressure under these con- 
ditions is practically impossible, but the difficulties can be 
partly overcome by very frequent firing of small quantities, 
so that the door is open a very short time and also that the 
layer of fuel is very thin and does not cut off much heat. 

4. The cleaning of the fire necessitates keeping the 
fire door open for several minutes, with results of the same 
variety as those just enumerated. 

Summing up these difficulties, they divide themselves 
into two classes — those which can be almost or entirely 
eliminated by skill of a very high order and those which are 
inherent and cannot be eliminated by skill. It will also 
be observed that all should give more trouble with fuels 
high in volatile than with those of the anthracite variety, 
both as to incomplete combustion and to the formation 
of smoke. 

Several other methods of hand firing have been proposed, 
particularly for use with bituminous coals, and some of 
them have been successfully utilized in isolated instances. 
Nearly all depend upon covering only part of the fuel bed 
at one time and, by alternating the parts covered in this 
way, fresh fuel on one part of the bed is coked while air is 
heated by coming in contact with the uncovered incandescent 
part of the bed and is therefore in proper condition to burn 
more perfectly the volume of hydrocarbons being distilled 
off. These methods are all good, but they involve a great 
deal of careful work and a high degree of skill on the part of 
the fireman. 

Other methods of eliminating some of the difficulties 
depend upon modifications of the furnace and air supply. 
Most attempt to entirely surround the fuel and the gases 
given off with heavy masses of brick work and tile, so that 



STEAM BOILERS 



335 



enough heat will be stored during incandescent periods 
to tide over the periods of cooling. Some forms have 
combined with this idea a series of air ducts in the brick 
work so arranged that air on its way to the furnace passes 
through these ducts and is heated. In some cases the air 
supply is automatically controlled and more air is supplied 
above the fire during the period of distillation, or coking, 
as it is called, than during the following period, when the 
coked coal is brightly incandescent and little volatile 
matter is present. 




Fig. 195. 



In some hand-fired furnaces which are intended for use 
with bituminous coals that give a long flame the parts of the 
boiling vessel within range of the flames are covered with 
tiles. This prevents impingement of unburned gases upon 
cool surfaces and thus tends to prevent the formation of 
smoke and incomplete combustion. 

Carrying this principle to its logical conclusion results 
in the installation of the grate in a firebrick chamber in 
front of the boiler-setting proper, as shown in Fig. 195. 
Such a device is known as a Dutch oven and is often very 
efficient in totally or partially preventing the formation of 



336 STEAM POWEE 

smoke. It does not, however, give as high an economy 
as might be expected, because a great part of the radiant 
heat of the fire does not reach the boiler surfaces and because 
the large external surface results in great radiation losses 
to atmosphere. 

Another interesting modification consists of reversing 
the direction of the draft, that is, the direction in which the 
air passes through the fuel bed. The type of furnace al- 
ready described is known as an updraft furnace, because the 
air passes upward in flowing through the bed. The modi- 
fied type here referred to is called a downdraft furnace, 
because the air flows downward in passing through the fuel. 

In downdraft furnaces the coal is fired on top of the 
grate as in other types, but the air is admitted above, flows 
downward toward what would normally be the ashpit, 
and from there on over the heating surfaces of the boiler. 
Fresh coal fired on top of the incandescent bed in such a 
furnace distills as in other types, but the volatiles are mixed 
with the entering air and are carried downward through the 
hot bed so that ideal conditions for combustion are more 
nearly attained. In some forms there is a second updraft 
grate beneath the downdraft grate. This second grate 
receives partly burned coals falling through from the upper 
grate and holds them until combustion is practically com- 
pleted. 

In downdraft furnaces the grate bars are generally made 
of pipes, and water, from the boiler or on its way to the 
boiler, is circulated through them. If this were not done 
the grates would quickly warp out of shape and ultimately 
burn away because of the high temperatures to which they 
are subjected. 

146. Mechanical Grates. In order to overcome the dif- 
ficulties arising from opening the doors for the purposes of 
cleaning the fire, numerous so-called rocking, shaking, 
self-cleaning, or dumping grates have been developed. 
These are generally built up of grate bars which have a 



STEAM BOILERS 



337 



rough T or an inverted L section with the upper horizontal 
branch of the T or inverted L slightly rounded,- as shown in 
Fig. 196. These bars are arranged in groups with their 
longitudinal axes running across the grate, and they are so 
supported that they can be rocked about a point in the verti- 
cal leg of the T or L by means of levers located at the front 
of the boiler. By rocking the bars the lower part of the fuel 
bed which has been burned to ash can be dropped into the 
ash pit, while the upper part is sufficiently agitated to close 
up holes which may have formed, and this can all be done 




Fig. 196. 



with the doors closed. Or, if desired, part or all of the fuel 
bed can be dropped into the ash pit by a similar rocking 
motion. 

147. Smoke and Its Prevention. An idea of the reasons 
for the formation of smoke will have been obtained from the 
preceding paragraphs. A reasonably skillful fireman should 
have little difficulty in burning anthracite coals in the simpler 
forms of furnaces without smoke, but it is almost impossible 
to commercially burn many of the varieties of bituminous 
coals in this way without the formation of excessive volumes 
of dense black smoke at intervals immediately following 
each firing. 



338 STEAM POWER 

Aside from all aesthetic and sanitary considerations, 
smoke is undesirable because it represents poor furnace 
conditions and waste. The actual loss of carbon in visible 
smoke is generally almost negligible in comparison with 
the other losses in the form of unburned hydrocarbons, 
the lowered initial temperature, etc. All of these losses 
combined represent a waste of considerable magnitude. 

The proper method of smoke elimination is not the 
combustion or removal of smoke already formed, but it 
is the burning of fuels in such ways as not to form any 
appreciable quantity in the first place. To accomplish 
this end the following must be achieved: 

1. Coal must be fired continuously and uniformly 
without the opening of doors which admit cold air to the 
furnace. 

2. Volatiles must be distilled continuously and uni- 
formly and in such a place that they are given ample oppor- 
tunity to mix with proper proportions of air and to burn 
completely before coming in contact with cool surfaces. 

3. The air supply must be properly controlled and 
tempered to meet the demands of the fuel both in and 
above the bed. 

4. The fire bed must be worked continuously and 
uniformly so as to eliminate ashes as rapidly as formed 
and to maintain a bed of uniform depth and condition. 

Some of these necessary conditions can be attained by 
the use of the various forms of hand-fired furnaces already 
described but, even in the hands of skillful and industrious 
men, it is impossible to meet all of them. Mechanical 
stokers which more nearly approach the ideals set have 
therefore been developed and are widely used. 

148. Mechanical Stokers. These mechanical devices are 
useful for two reasons — they eliminate a great deal of labor 
and they make possible the burning of many varieties of 
refractory fuels without the formation of excessive quanti- 
ties of smoke. 



STEAM BOILERS 339 

Despite the good results which can be achieved by their 
use, mechanical stokers are not installed in small plants 
as often as might be expected. This is because good stokers 
are very expensive in comparison with hand-fired furnaces 
and, despite economy of fuel, do not generally show a finan- 
cial saving unless their use eliminates the services of several 
firemen. 

It is generally assumed that one man can care for water, 
coal and ashes for about 200 boiler horse-power or can 
handle coal only for about 500 boiler horse-power. Experi- 
ence has shown that one man can fire about 2000 to 5000 
boiler horse-power when the boilers are equipped with 
good stokers and coal-handling apparatus. 

Financial calculations will not justify mechanical stokers 
in many of the smaller plants in which they are used. How- 
ever when improved conditions with respect to smoke and 
dirt and improved labor conditions are taken into account 
they are generally regarded as paying propositions. 

Mechanical stokers can be roughly divided into two 
types, those which duplicate hand spreading of fuel and 
are known as sprinkler stokers, and those which supply 
fuel at one or more points and work it progressively toward 
the ash end of the apparatus as it burns. The first type 
has not been widely installed, though it is possible that 
it may meet with more popular approval after further 
development. 

Stokers of the second type may be roughly divided 
into four classes, which are 

1. Chain grates. 

2. Inclined stokers or overfeed stokers. 

3. Underfeed stokers. 

4. Combinations of above. 

A chain grate, as made by the Illinois Stoker Company, 
is illustrated in Figs. 197, 198, 199, and 200. It consists 



340 



STEAM POWER 




O 



STEAM BOILERS 



341 



of a broad chain made up of a great number of small links 
and carried on toothed wheels and roller wheels supported 
in a frame which can be wheeled into position within the 




Fig. 198. — Sprocket and Links of Illinois Chain Grite. 





TOP VIEW OF CHAIN 

SHOWING DISTRIBUTION OF AIR SPACES 



BOTTOM VIEW OF CHAIN 

SHOWING ROLLERS FOR DRIVING-SPROCKET 
ENGAGEMENT 



Fig. 199. 

boiler setting. The general arrangement of the chain and 
rollers is shown in Fig. 197; details of the front or driving 
rollers and of the links are shown in Fig. 198; a top and bot- 
tom view of part of the chain is given in Fig. 199; and 
Fig, 200 is a perspective view of the frame showing the 



342 STEAM POWER 

tracks on which it may be rolled into and out of the boiler 
setting. 

The chain is driven slowly in the direction indicated 
by the arrows in Fig. 197 by power applied, through worm 
gearing, to the shaft of the toothed wheels at the front 
of the stoker. Coal feeds automatically from the hopper 
by gravity and is carried into the combustion space by 
the moving chain, the thickness of the bed being controlled 




Fig. 200.— Framework of Illinois Chain Grate. 

by the height of the adjustable gate shown. As the fuel 
enters the furnace it passes under the coking arch, which 
spans the entire front part of the grate and which is main- 
tained at a high temperature by heat radiated from the in- 
candescent fuel nearer the inner end of the grate. The 
volatiles are distilled from the fresh coal by heat received 
from this arch and are heated and mixed with air at this 
point. The coked fuel is then carried on into the furnace 
and burned, the refuse being discharged at the bridge wall. 
If the thickness of bed and speed of chain travel are 
properly adjusted, all of the fuel can be coked before pass- 
ing out from under the arch and can be burned almost 



STEAM BOILERS 



343 



completely before reaching the bridge wall, so that prac- 
tically ashes only will be discharged. 




The apron shown at A in Fig. 197 is used to prevent 
the free passage of air to the part of the chain carrying 



344 STEAM POWER 

practically nothing but ash, as this would result in excessive 
dilution of the products of combustion. 

A stoker of this type installed under a horizontal return- 
tubular boiler is shown in Fig. 201. In the illustration part 
of the side frame of the stoker is broken away in order to 
show the chain and its roller guides. The eccentric shown 
near the top of the front of the boiler drives the chain 
through an arm of adjustable length, which makes possible 
the control of the speed of chain travel. 

The early forms of chain grates were intended for use 
with natural draft, that is the space under the grate con- 
nected directly with the atmosphere and maintenance of 
a " draft " or " under-pressure " in the furnace permitted 
the external atmosphere to push through the fuel bed, this 
supplying the necessary oxygen for combustion. As the 
art developed a demand arose for stokers capable of burn- 
ing more fuel per square foot of grate area. This demand 
was met by developing stokers which could be used with 
" forced draft." These stokers were so arranged that air 
could be forced through the fuel bed by fans discharging 
into the space beneath the stoker. 

The better designs of forced draft, chain grates provide 
for differential air supply to different sections of the fuel 
bed. For this purpose the space between the upper and 
lower chains is divided into compartments by vertical 
partitions in planes at right angles to the travel of the grate. 
Provision is then made for controlling the air pressure in 
each box separately by means of dampers or equivalent. 
With such construction the air pressure applied to the under 
surface of the grate can be regulated by sections instead of 
for the entire grate and it becomes possible to grade the 
air supply to more nearly meet the requirements of the 
fuel in different stages of combustion. 

As an example of the usefulness of such an arrangement, 
consider the last section over which the material passes 
before being discharged from the end of the stoker. If 



STEAM BOILERS 345 

the fuel is very nearly burned out before it reaches this 
section a very small amount of air will suffice for the com- 
pletion of combustion. Any greater amount would simply 
pass through without useful effect, would serve to lower 
the furnace temperature and to increase the excess coeffi- 
cient. On the other hand if the fuel still contains large 
amounts of unburned carbon when it reaches this section 
it is desirable to be able to supply a relatively large amount 
of air so as to more nearly approximate complete com- 
bustion of the carbon before the material is discharged 
from the grate. 

Development of the forced draft chain grate and of 
numerous small but important refinements in details of 
design has brought the chain grate into particular promi- 
nence in connection with the combustion of certain very poor 
grades of fuel. Thus very fine sizes of anthracite previously 
considered of no commercial value because of their small size 
and high ash and moisture content are successfully burned 
on chain grates particularly designed for such use. Some 
of the poorer varieties of bituminous fuel with high ash 
content of sux?h character as to form excessive amounts of 
clinker, are also being burned on such devices. 

An inclined stoker with front feed and a step grate, 
known as the Roney stoker, is shown in Figs. 202 and 203. 
The fuel is fed out of the hopper and onto the dead plate 
by means of the reciprocating pusher. From the dead 
plate it is pushed down upon the grate bars by the follow- 
ing fuel. These bars are rocked mechanically so that their 
tops alternately assume horizontal and inclined positions, 
and this action feeds the fuel downward until it is dis- 
charged onto the dumping grate. The material collect- 
ing on this grate is periodically dropped by hand into the 
ashpit. 

The fuel is coked while passing under the coking arch 
and the coked material is practically completely burned 
by the time it has traveled down the grate, The volatile^ 



346 



STEAM POWER 



are mixed under the coking arch with heated air which has 
passed through the grate and with heated air forced in 
above the fuel. 

An inclined stoker of the side-fees type with bar grates 
known as the Murphy stoker, is illustrated in Figs. 204, 
205 and 206, This stoker is provided with two coal- 




Sheath 

Agitator 

Lock-Nut 



Connecting-Rod 



Fig. 202. — Details of Feed Mechanism, Roney Stoker. 



magazines or hoppers which are placed horizontally in 
the side walls of the boiler setting and feed fuel onto the 
inclined grate bars, Fig. 204, which carry it downward 
toward the lower point of the V formed by the grates. 
The grate bars, Fig. 206, are alternately fixed and mov- 
able, the movable bars being hung from above and their 
lower ends being moved up and down by power furnished 
by a small steam engine or other convenient source. 



STEAM BOILERS 



347 




m 



3 



348 



STEAM POWER 



A toothed bar arranged for rotation by hand or by 
power is located at the bottom of the V and is used for 
grinding up ash and clinker which is too large to fall through 
into the ash pit. This bar is kept cool by making it hollow 
and connecting one end to the smoke flues or stack so that 
air is constantly drawn through it. 



iP£ 




^ 

£/&■;, 



Fig. 204. — Transverse Section of the Murphy Stoker. 



The location of the coking arch and the method used 
for supplying warm air should be evident from the figures. 

A stoker of this variety is shown in place under a hori- 
zontal water-tube boiler in Fig. 207. 

An underfeed stoker made by the Combustion Engineer- 
ing Company is shown in Figs. 208 and 209. Coal is fed 
from the hopper onto the reciprocating bottom B by means 
of the reciprocating pusher P. The part B forms the bottom 
of a trough as shown in Fig. 209, and its reciprocating 
motion feeds the coal upward and out of this trough so that 



STEAM BOILERS 



349 



it spills over onto the inclined grate bars. The reciprocating 
motions are all obtained from the direct-acting steam cylin- 
der shown. 

The inclined grate bars are alternately fixed and mov- 
able, the movable bars sliding back and forth at right angles 



W^ZZ ZW A.-rL*.- ■»: J 




Fig. 205. — Longitudinal Section, Murphy Stoker. 



to the trough under the action of horizontal rocking bars 
R. This action gradually feeds the fuel downward and 
toward the side of the furnace, the refuse finally landing 
on the dumping trays shown. 

Air enters the duct below the trough through the 
adjustable gate G, controlled by crank C, and part of it 
passes out through holes H hear the top of the trough, 



350 



STEAM POWER 



Fig. 209. The remainder passes down through the hollow 
grate bars and into the heated air box from which it flows 
upward between the grate bars. 

It will be observed that the coal is fed onto the grate 
from below, so that all volatiles distilled off must pass up- 
ward through the incandescent fuel before entering the 
space above the fuel bed. Part of the air which is to burn 



Stationary 
Grate Bar 




Movable 
Grate Bar 



Fig. 206.— Grate Bars of Murphy Stoker. 



this volatile matter also passes through the fuel bed and 
the remainder flows over the incandescent fuel from the 
opening shown near the hopper in Fig. 208. The air and 
volatiles are thus raised to a high temperature and well 
mixed, and the operation is continuous and uniform, all 
tending to facilitate smokeless combustion. 

Another variety of underfeed stoker known as the 
Taylor stoker is shown in Fig. 210 (a), (b) and (c). 
This stoker is built up of alternated retorts and air 



STEAM BOILERS 



351 



boxes, the proper number to give the desired width 
of stoker being used. Coal is fed from the hoppers 




W 






into the retorts by the upper ram or plunger shown 
in Fig. 210 (b) and part of it is again pushed forward by 



352 



STEAM POWER 




Steam 
Cylinder 



Fig. 208.— Longitudinal Section of Type "E" Stoker 







Fig. 209.— Cross Section of Type "E" Stoker. 



STEAM BOILERS 



353 




354 STEAM POWER 

the lower ram or plunger. The stroke of the lower 
plunger can be regulated and in this way the relative 
quantities of coal pushed forward in the upper and lower 
parts of the retorts can be controlled. The coal spreads 
over the tuyere blocks which form the inclined tops of 
the air boxes and forms a comparatively even, inclined 
layer of fuel. 

Coking proceeds under the incandescent fuel which 
forms the upper surface of this layer, and the volatiles 
mix with air entering through the hot tuyeres and pass 
upward through the hot fuel above. 

In this stoker advantage is often taken of the fact that 
the draft (pressure of air) required with underfeed stokers 
is so great that it can be more economically attained by the 
use of a fan than by the use of a stack. The fan and the 
coal-feeding plungers are both connected to one engine 
and the speed of this engine is automatically controlled by 
the steam pressure within the boiler. As this pressure 
decreases the engine speeds up, thus delivering more coal 
and air and as the pressure increases the engine slows down 
with opposite results. By properly fixing the travel of the 
plungers initially, the best relative proportions of air and 
coal are set for the entire range of loads to be carried and the 
variation of both is thereafter in approximately the same 
proportions. 

A stoker of this type in position under a horizontal 
water-tube boiler is shown in Fig. 211. A double-ended 
arrangement of Taylor stokers as used under very large 
water-tube boilers is shown in Fig. 212. 

Powdered or pulverized coal burning equipments have 
been invented in great numbers and are successfully used 
in several of the industries. They are essentially stokers. 
Many attempts to use pulverized coal for firing boilers have 
been made from time to time but until quite recently the 
results were not considered satisfactory. Within the past 
few years a number of installations have been made in 



STEAM BOILERS 



355 



boiler plants and the results are claimed to be satisfactory 
in many instances. 

The coal after crushing to moderate size is dried by 
passing it through a drier heated by hot products of com- 




Fig. 211. — Taylor Stoker Under Horizontal Water-tube Boiler. 



bustion unless the moisture content " as received " is not 
considered too great for satisfactory pulverization and utiliz- 
ation. The material is then pulverized in a mill arranged 
to give very fine subdivision, pulverizing so that at least 
60 to 90 per cent passes through a 200 mesh sieve being 
common practice. 



356 



STEAM POWER 




Fig. 212. — Double-ended Arrangement of Taylor Stoker under Sterling 
Type W. Boiler. 



STEAM BOILERS 357 

The pulverized fuel is then transported to small hop- 
pers immediately adjacent to the individual boilers. From 
these hoppers it is fed into a stream of air in which it 
becomes suspended and by which it is carried through 
the burner into the furnace. The burner generally consists 
of a simple nozzle. The fuel burns in the form of a torch 
at the end of the nozzle, combining with the air which 
carried it in and generally with additional air admitted 
through controllable doors or openings in the walls of the 
furnace. 

The fine pulverization makes possible very rapid com- 
bustion of the individual particles and the flames have the 
appearance of an intensely hot gas flame or liquid fuel 
flame. Success depends on completing combustion while 
the particles of fuel are still suspended in space, that is 
before they have had time to settle or to impinge on any 
solid surface. The flame is easily brought to so high a 
temperature that the ash contained in the fuel is liquefied. 
Unless the installation is carefully planned and properly 
operated this fused ash is very apt to cause serious difficulties 
by building up stalagmites and stalactites in the furnace 
or by actually solidifying on the cold heating surfaces of the 
boiler. 

Many advantages are claimed for pulverized fuel firing 
in boiler plants. The principal claims are great flexibility 
with respect to character and quality of fuel burned, ease of 
operation so that one fireman can handle a large boiler 
installation, ease and simplicity of regulation so that high 
thermal efficiency can be obtained and maintained even 
under rapidly changing conditions, and elimination of 
banking losses as fuel is entirely excluded from the 
furnace when steam is not desired. However, the use of 
pulverized fuel entails a large investment in equipment 
for preparing it and the fixed charges on this equipment 
as well as the operating expenses chargeable to prepara- 
tion tend to balance gains resulting from the advantages 



358 STEAM POWER 

enumerated above. The art is as yet so young that it 
is impossible to obtain sufficient data to make a true 
and complete comparison between pulverized and solid 
fuel firing. 

Oil firing is essentially a mechanical, rather than a 
manual process, and while oil burners are not ordinarily 
understood as belonging to the class of mechanical stokers, 
they have all the essential characteristics of such ap- 
paratus. 

To burn oil successfully under a boiler it must be 
finely atomized and mixed with the necessary quantity 
of air, and there must be sufficient open space within 
the furnace for the free development of the flame and 
the completion of combustion before impingement on cool 
surfaces. 

Oil-burning furnaces are generally given a rather large 
volume; considerable firebrick is used in such ways as to 
give incandescent walls and baffles to assist ignition and 
combustion, and all heating surfaces are arranged so that 
they are not in the direct path of the flame. 

The atomization of the oil is effected in two distinctly 
different ways. In some forms of burners it is brought 
about by mechanical means, the oil being pumped through 
a nozzle of some sort which is so shaped that the issuing 
jet breaks up into a great number of very small particles. 
In other forms, steam is used to break up the jet, the steam 
and oil entering the body of the burner separately and later 
coming into contact in such a way that the oil is literally 
torn apart by the steam. This form of burner has been 
more extensively used in the United States than has the 
former but present developments indicate a probable 
reversal in this respect. The mechanical type of burner 
has been developed to a very high degree during the past 
few years, and one form, at least, seems to possess marked 
advantages. 

Oil burning shares with the burning of powdered coal, 



STEAM BOILERS 350 

the property of permitting very accurate regulation of 
the air supply to suit the quantity of fuel being burned. 
The excess coefficient may therefore be maintained at a 
low value and the initial temperature may be made cor- 
respondingly high. Part of the advantage thus gained 
over the commoner methods of coal firing is, however, 
counterbalanced by the quantity of steam used for heat- 
ing and pumping the oil and for atomizing in some forms 
of burners. 

Both oil burning and powdered-coal burning can be 
easily made to give smokeless combustion in properly 
designed furnaces and both yield readily to forcing. That 
is, the temporary consumption of excessive quantities of 
fuel to tide over short demands for excessive amounts of 
steam is comparatively easily effected if sufficient furnace 
volume is available. 

149. Rate of Combustion. The rate at which coal is 
burned in a given furnace or on a certain grate is generally 
given in terms of pounds of coal fired per square foot of grate 
surface per hour and is referred to as the rate of combus- 
tion. 

The rate at which coal can be consumed is largely 
dependent on the intensity of draft available, that is, on 
the air pressure available for driving air through and 
over the bed of fuel. The higher the pressure available, 
the greater will be the quantity of air which can be sup- 
plied and the greater will be the quantity of coal that 
can be burned. If it were not for the cost of creating 
the draft, the only limit to increasing the rate of com- 
bustion would occur when the velocity of the air became 
so great that the fuel would be picked up from the grate 
and carried onward into the flues in a partly burned con- 
dition. Commercial drafts give pressure differences above 
and below the fuel bed which range from about 0.1 inch 
of water to as high as 8 ins. In stationary plants the 
pressures generally range from 0.1 to about 0.5 in cases 



360 STE^M POWER 

where hand firing is employed, and are carried as high as 
5 or more inehes of water with some forms of mechanical 
stokers. 

The best rate of combustion varies with the type and 
size of fuel, the type and size of furnace, the type and size 
of boiler, the draft and many other considerations. In 
ordinary power-plant practice with hand firing the rates 
of combustion commercially used generally fall within the 
following limits: with anthracite, 15 to 20 lbs. per square 
foot per hour; with semi-bituminous, 18 to 22 lbs.; and 
with bituminous, 24 to 32 lbs. In the case of stokers, these 
values may be doubled and even trebled if proper provisions 
are made. 

As practically all of the volatile is consumed above 
the grate, the fixed carbon content is practically the de- 
termining factor, since it is this constituent that is 
.burned on the grate. This explains the high rate possible 
with fuels with high volatile content. The most eco- 
nomical results are generally obtained when from 12 to 
16 lbs. of fixed carbon are consumed per square foot of 
grate per hour. 

The figures given above do not represent limiting con- 
ditions. In torpedo-boat practice, where high-draft pres- 
sures are used (from 4 to 8 ins. of water), rates of from 
50 to 120 lbs. are attained. On locomotives, which also use 
high-draft pressures, rates of combustion greatly in excess 
of stationary practice are generally used. 

The capacity of a given boiler, that is, its ability to 
generate steam, increases as the rate of combustion is 
increased, since more heat is thus made available. The 
economy of the combination, that is, pounds of steam 
generated per pound of coal fired, increases until some best 
rate of combustion for the fuel in question is reached, and 
thereafter decreases. The variation of economy is, however , 
not very great for a comparatively wide range of combustion 
on either side of the best rate. 



STEAM BOILERS 



361 



Curves giving approximate draft pressures required fee 
different rates of combustion when different kinds and sizes 
of fuel are hand fired are given in Fig. 213. The sizes 
referred to are explained in Tables XIII and XIV. Table 
XIII also shows the relative increase of ash content as the 



a? ° • 














/ 


/ 




/ 
















/ 














t 


j/ 


















5 

V 


4 






^5^^ 


V 






V 




V 


..^ 












I 






f^ 




ie 














— ^Ui 


&«»•-' 









5 10 15 20 25 30 35 40 45 60 

Pounds of Coal per Sq.. Ft. of Grate Surface per Hour 

Fig. 213. — Draft Required for Different Rates of Combustion with 
Different Sizes and Kinds of Fuel. 



size decreases, there being a tendency toward the concen- 
tration of the ash in the smaller sizes. 

150. Strength and Safety of Boiler. Attention has 
already been called to the fact that the boiling vessel has to 
be designed with two different requirements in view: it 
must be mechanically strong to resist internal pressure and 
it must transmit the maximum amount of heat to the con- 
tained water. 

Spherical and cylindrical surfaces with the pressure act- 
ing on the inside of the curve are best adapted to resist 
such pressures, as they already have the shape which the 
pressure would tend to give them. Boilers are, therefore, 



362 



STEAM POWER 



TABLE XIII 

Sizes of Anthracite Coal 

(Sizes larger than pea coal generally too costly for power-plant use.) 



Name. 


Through Screen 
with Mesh. 
(Inclusive.) 


Over Screen 
with Mesh. 
(Inclusive.) 


Ash Content 
(Average). 


Run of mine 

Broken. 


unscreened 


unscreened 
21 
2 

li 

3 
4 

1 
2 
1 
4 
1 
8 




Egg 

Stove 

Chestnut 


2| 

2 

li 

3 

4 
1 
2 
1 
4 


6 
10 
13 


Pea 

Buckwheat No. 1 

Buckwheat No. 2 or rice. . . 


15 
17 

18 



TABLE XIV 

Sizes of Bituminous Coal 

(Considerable variation in commercial practice in £.~~~ng and sizing.) 



Name. 


Through Bars 

Spaced Apart. 

(Inches.) 


Over Bars Spaced 

Apart. 

(Inches.) 


Lump 




li 


Nut 

Slack 


li 

4 


4 



constructed as far as possible of vessels having only spherical 
and cylindrical surfaces. 

"Flat surfaces which are poorly adapted to resist such 
pressures as act within a boiler must often be used despite 
their weakness. When incorporated in a boiler they are 
invariably " stayed," that is, braced by being fastened to 
other surfaces by stay bolts and other forms of fastenings. 
Examples will be given later. 

Most of the early designs of boilers and many of the 
modern types consist of large cylindrical vessels made 
by riveting together properly shaped steel plates. These 
shells are often traversed from end to end by flues or 



STEAM BOILERS 



363 



tubes for carrying hot gases and generally have flat ends 
more or less perfectly braced by these tubes and by long 
tie rods and other braces. Such boilers when in operation 





Fig. 214. — Lap Joint. 



Fig. 2 15.— Butt Strap Joint. 



are almost entirely filled with water and often hold many 
tons. 

Boilers of these types have been responsible for many 
disastrous boiler explosions, and this fact has led inventors 
to the development of models which should be less dangerous. 
It seems practically impossible to develop a commercial 
boiler which cannot be made to explode to a certain extent 





Fig. 216.— Riveted Plates of Boiler Shell. 



Fig. 217. 



if sufficiently mistreated and mishandled, but much can be 
done to minimize the danger. 

The great weakness of the older forms lies in the riveted 
joints, which can never be made as strong as the plates 
which they fasten together. Two types of joint are in use; 
they are known respectively as the lap joint and the butt 
strap joint. These are shown in Figs. 214, 215 and 216. 
So far as a circumferential seam, that is, one running around 
the cylinder as shown in Fig. 216, is concerned, the lap joint 



364 STEAM POWER 

is perfectly satisfactory and is universally used. With 
longitudinal seams, however, this is not the case. A lap 
joint throws the joined edges out of a true cylindrical sur- 
face as shown in Fig. 217, and when the vessel is subjected 
to pressure there will be a tendency for the plates to assume 
a cylindrical contour as nearly as possible. This causes 
local bending of the plates on each side of the lines of rivets, 
and the continued repetition of this action ultimately causes 
failure. The conditions are often made still worse by calk- 
ing the joint on a line indicated by a in Fig. 217, that is, 
by hammering the metal at the inner surface of the edge 
of the outer plate into firmer contact with the outer sur- 
face of the inner plate for the purpose of making a tight 
joint. 

The butt-strap joint can obviously be made so that 
the joined plates more nearly form a true cylindrical sur- 
face. 

Other weaknesses of the older forms lie in the flat surfaces 
used; in constructions which render it possible for sediment 
to collect on heated surfaces and thus permit local over- 
heating of the plate; and, above all, in the very large 
quantity of water contained. 

The disastrous consequences of boiler explosions are 
generally due to the action of the hot water contained within 
the boiler and not to the steam contained at the time rupture 
occurs. The water within the boiler is under steam pressure 
and approximately at steam temperature. Removal of the 
pressure by rupture of the container would enable a great 
part of this water to flash suddenly into steam at the 
expense of its own heat, and this is exactly what occurs in 
the case of a boiler explosion. Local failure causes a sudden 
lowering of pressure, and this results in the formation of 
large volumes of steam which, blowing out through the 
initial fracture, tend to enlarge it, to move the boiler and 
surroundings, and, in general, to do all possible to further 
the rupture and make conditions worse. 



STEAM BOILEES 365 

From the preceding discussion the requirements for 
maximum safety can be deduced. They are: 

1. The smallest convenient diameter of cylindrical ves- 
sels, so as to decrease the total load on joints for any 
given steam pressure. 

2. The elimination of the greatest possible number of 
riveted joints and the use of butt-strap longitudinal joints 
on all large-diameter, cylindrical vessels. 

3. The substitution of curved surfaces for all flat stayed 
surfaces. 

4. So shaping the boiler that the required extent of 
heating surface may be obtained without enclosing a great 
volume to be filled with hot water when the boiler is 
steaming. 

5. So shaping the boiler that such water as is contained 
therein will be divided up into small masses contained 
within separate vessels connected in such a way that rapid 
flow of all water toward one point of failure is impossible. 

6. So shaping the boiler that no riveted joints shall be 
in the paths of flames and that no sediment can collect on 
metal immediately over flames or exposed to very hot gases. 

7. So shaping the boiler that it shall 
be free to expand and contract with 
changes of temperature, with the least 
resultant strain on the different parts. 

These various requirements are most 
nearly met in the different forms of water- 
tube boilers, some of which will be de- 
scribed in succeeding paragraphs. 

151. Circulation in Boilers. If a flask 

of water, such as that shown in Fig. 218, */ G : 218 '~ i? ir( : u ~ 
. , . , . ,, . ,. . , ,' lation in a r lask. 

be heated in the manner indicated, the 

water will gradually acquire motion and follow paths 

such as those shown by the arrows in the illustration. The 

heated water will rise in the center of the mass and the 

cooler water will flow downward around the outer surface. 




366 



STEAM POWER 



Such motion is called circulation. Rapid circulation within 
a boiler is very desirable, since it brings the maximum quan- 
tity of water in contact with the heating surfaces in a given 
time and hence tends to increase the amount of heat taken 
from those surfaces. It also tends to sweep along any 
bubbles of steam or gas formed on such surfaces and to carry 
away any sediment which may have collected, thus pre- 
venting overheating of the surfaces. 

Circulation can be expedited by providing free and 
unrestricted paths for the water so as to guide it in the 
proper directions and by applying the most intense heat 
at the proper point along the path of the water. The tem- 
perature of the water which is subjected to the most intense 



mLbhki^ 



~m~- 



^=^iM ^= 



& 




Blo-w oil' 

Fig. 219. — Elementary Types of Boilers. 



heat is naturally raised and the water at that point becomes 
less dense than in other parts of the boiler. The formation 
of steam at such points also materially lessens the density. 
As a result of this lowering of density the heated water 
rises and the cooler water descends to take its place. The 
more rapid this exchange can be made, the more steam can 
be generated from a given amount of surface in a given 
time and hence, other things equal, the better the boiler. 

The elements of two common forms of boiler are shown 
in Fig. 219, the arrows indicating the direction of the cir- 
culation and its effect upon the delivery of steam and of 
sediment. 

152. Types of Boilers. In a book of this scope it would 
be impossible to describe all the types of boilers at present 



STEAM BOILERS 



367 



in use. The more important varieties have therefore been 
chosen for description and illustration. 




Two types of internally fired, tubular boilers more 
accurately described as internally fired, upright or vertical, 
fire-tube boilers are shown in Fig, 220. The furnace is 



368 



STEAM POWER 



Tube Sheet 

Steam Space 

Exposed Tubes 

Water Level 



Water Column 
and Try Cocks 



Feed Water 
Connection 




Pressure 
Gauge 



Hand Holes 

(Closed by 
hand hole covers 
when in operation) 



Fig. 221.— Large Internally Fired Tubular Boiler. 



STEAM BOILERS 369 

contained within the shell of the boiler and is almost com- 
pletely surrounded with water. The heat radiated from 
the hot fuel is thus almost entirely received by the water 
of the boiler. The hot gases, rising from the fuel bed, 
pass upward through the tubes and, after giving up part 
of their heat to the surrounding metal, enter the smoke 
box and pass directly to the stack. 

Boilers of the type shown in Fig. 220, (a) and (b) are 
called exposed-tube boilers, because the water level is carried 
below the tops of the tubes. The tubes, therefore, extend 
through the steam space and act as imperfect superheaters. 
Boilers of the type shown in Fig. 220, (c), in which the tubes 
do not enter the steam space, but are entirely covered by 
water, are called submerged-tube boilers. 

Upright tubular boilers of the types shown in Fig. 220 
are built by a number of manufacturers in sizes ranging 
from about 4 boiler horse-power to about 50 boiler horse- 
power. They are self contained, require no setting of any 
kind, and are shipped completely erected. Such boilers 
are very often mounted on trucks or skids and used to 
generate steam for small hoisting and other forms of con- 
tractors' engines. They are also used on steam fire engines. 

The pressure carried in these small tubular boilers is 
generally under 100 lbs. per square inch, but they can be 
built for higher pressures if desired. 

In Fig. 221 is shown a larger type of internally fired 
tubular boiler as made by the Bigelow Company for station- 
ary use. These boilers are similar to those just described, 
but are made only in large sizes, in this case, in sizes ranging 
from 40 boiler horse-power to 200 boiler horse-power. The 
exposed tubes generally give a superheat of about 25° F. 

These large upright boilers can be built to operate with 
a pressure as high as 200 lbs. per square inch and because 
of the small area covered by even the largest sizes, they 
are particularly adapted to locations in which floor space 
is limited. 



370 



STEAM POWER 



The locomotive type of boiler is shown in Fig. 222. It is 
an internally fired, horizontal, tubular or fire-tube, boiler. 

Such boilers are seldom used for stationary purposes, 
but are universally used on steam locomotives and, in the 
smaller sizes, are often mounted on trucks or skids and used 
for semi-stationary purposes by contractors and others. 
Boilers of this type are built in sizes ranging from 10 boiler 
horse-power or less up to over 100 boiler horse-power for 
general power purposes, while those used on the largest 
locomotives generate over 2000 boiler horse-power. 

The Continental type of boiler, named from the Con- 
tinental Iron Works, is shown in Fig. 223. These boilers 







Handholes 

Fig. 222. — Locomotive Type of Boiler. 



may be described as internally fired, return tubular, with 
semi-external combustion chamber, this chamber being out- 
side of the boiler shell proper but being built as an integral 
part of the boiler and transportable therewith. Boilers cf 
this type are built in sizes ranging from about 75 boiler 
horse-power to 300 or more. 

The grates, furnace and ash spaces, and bridge wall are 
all carried within circular, corrugated flues, one flue being 
used in the smaller sizes and two in the larger. The corru- 
gations serve the double purpose of strengthening the flue 
and of exposing added heating surface to fire and hot gases. 

The steam pipe shown just below the steam connection 
at the top of the boiler is commonly used on boilers for the 



STEAM BOILERS 



371 



paaiog noipog dox 







372 



STEAM POWER 



purpose of preventing the escape of excessive quantities of 
moisture with the steam. 

These boilers are very compact in shape and are short 
for their capacity, but they contain a great volume of water. 
They possess the advantages of having a large steam space 
and a very extended liberating surface over which the steam 
separates from the water. 



Uptake 




Tubes 



Fire Doors 



Man Hole- 

Fig. 224. — Scotch Marine Type Boiler. 

The Scotch marine type of boiler is shown in Fig. 224. 
It has the same general construction as that just described 
excepting that the combustion chamber is entirely enclosed 
within the water space of the boiler. This chamber is 
built up of flat plates and is held against collapse by numer- 
ous stay bolts. Boilers of this type were until recently 
the standard for marine practice, but they are now being 
replaced in many instances by water-tube boilers of more 
recent design. 



STEAM BOILERS 



373 



Scotcn marine boilers are very economical in the use of 
fuel, are good steamers, and are absolutely self contained. 
They are built in numerous sizes, the smallest having shells 
with diameters of about 6 ft., while the largest diameter 
used is about 16 ft. The largest boilers have three and 
four corrugated furnaces. 

Two types of externally fired, return-tubular (or 
" H.R.T.") boilers are shown in Figs. 225 and 226. The 




Fig. 225.— Horizontal Return-tubular Boiler with " Full Flush Front." 



only essential differences in these two types are in the forms 
of setting and in the methods of suspending the boilers. 
The shell is generally rigidly supported at the furnace end 
and arrangements made to allow for movement of the other 
end with changes of temperature. 

These boilers can be built very cheaply and are therefore 
widely used when their limitations, do not prevent. It has 
been found inadvisable to build them in sizes larger than 
200 boiler horse-power or for pressures higher than 150 
lbs. per square inch, and they are generally used in smaller 



374 



STEAM POWER 




STEAM BOILERS 



375 




376 



STEAM POWER 



Fig. 228.— Forged 
Header for Bab- 
cock & Wilcox 
Boiler. 



sizes and with lower pressures. These limitations are set 
by permissible thickness of metal immediately above the 
fire, experience having shown that the 
plates deteriorate rapidly at this point if 
made too thick. 

One form of Babcock & Wilcox water- 
tube boiler is shown in Fig. 227. This 
boiler is built up of sections consisting of 
several tubes joined at the ends by headers, 
and the sections are connected side by side 
at each end to a long horizontal drum. 
The ends of this drum are closed with 
" dished " heads, thus doing away with 
flat surfaces and the necessity for stays 
within the drum. 

A detail of the forged header is shown in 
Fig. 228. It may be regarded as a long 
box of rectangular section with opposite 
walls pierced by circular holes, which has been so distorted 
as to give it a wavy shape. The distortion brings the holes 
into such positions that the tubes when expanded into these 
holes are " staggered," that is, 
do not lie one above the other. 
The general principle in- 
volved in the arrangement of 
these sections or elements and 
the resulting circulation are 
shown in Fig. 229. The location 
of the feed- water inlet and other 
details are shown in Fig. 230. 
It will be observed that the feed 
water enters in such a direction 
and position that it is readily 
picked up by the current of water circulating in the boiler, 
carried toward the rear and down the rear header. During 
this travel it is heated by contact with the hot water in 




Fig. 229. — Elementary Babcock 
& Wilcox Boiler, Showing 
Circulation. 



STEAM BOILERS 



377 




Hand hole opposite end 
of tube, closed by hand 
hole cover when in 
operation. 



End of tube 
expanded into 
bole of header. 



Fig. 230. — Details of Babcock & Wilcox Boiler Construction, 



378 



STEAM POWER 



the boiler and most of its impurities are separated out and 
settle in the mud drum at the bottom of the rear header. 




The boiler is suspended by stirrups from beams carried 
by the brickwork as shown in Fig. 227, the tube sections 



STEAM BOILERS 



379 



simply hanging from the drum by the nipples at each end. 
The various parts of the structure are thus free to expand 
and contract independently as their temperatures change 
and are not bound in any way by the brick setting. 

The steam is collected from a perforated steam pipe near 
the top of the steam space. The baffle shown in Fig. 230 
prevents the steam which rises from the front header from 
carrying the water bodily into the steam space and makes 
the greater part of the water surface in the drum act as 
separating surface. 

The scale which accumulates inside of the tube is removed 
by tools inserted through the hand holes in the front headers 
opposite the ends of the tubes. One of these hand holes 
^nd its cover are shown in section in Fig. 230. Soot and 
dust which accumulate on the outer surfaces of the tubes 
are blown off periodically by a steam jet, the necessary 
nozzle and hose being in- 
serted through the tall and 
narrow side cleaning doors 
shown in Fig. 227 opposite 
each " pass." 

A section of the Heine 
water-tube boiler is shown 
in Fig. 231. This boiler con- 
sists of a slightly inclined 
drum with dished heads, two 
sheet-steel headers and nu- 
merous tubes connecting 
these headers. The shape 
of the header is shown in 
Fig. 232, which indicates the 
positions occupied by the tubes and the way in which 
the header is joined to the drum. 

The products of combustion are generally made to pass 
along the tubes by the longitudinal baffles shown, instead 
of across the tubes as in the boiler last described. 




o o o o o o o o o o o 

0°0°0 0°0°0°0°OoO°000 

o o o o o o o o o o o 

o°o°o°o°o°o°o°o°o°o°o°o 

o o o oooooooo 

O o o° O o o c o «• O o O o 0»0«OoO»0 
0<D<D OOOOOOOO 

O o o o o o o ° O o o ° o ° o *■ o ° o « o - o 
o o o o o o o o o o o 

coooooooo o°o o°o 



Fig. 232.— Front End Elevation, 
Heine Boiler. 



380 STEAM POWER 

The mud drum in this type is located within the boiler 
and consists of a sheet-steel box supported a few inches 
above the bottom of the drum. The feed water enters at 
the front end of this drum and gradually spreads out as it 
is heated by the surrounding water. The greater part of 
the impurities settles to the bottom and is blown off period- 
ically. The warmed water rises and flows out of an opening 
in the top of the box at the front end and there joins the 
circulation of the boiler, traveling toward the rear, down 
the rear header and up the tubes to the front header. 

The interior of the tubes is cleaned of scale through 
hand holes just as in the last boiler. The external surfaces 
are freed of soot and dust by means of a steam jet which 
is introduced through the stay bolts in the headers, these 
bolts being made hollow for this purpose. Since it is not 
necessary to use doors in the side walls for cleaning in this 
type, Heine boilers are often set up in batteries of three or 
more, each interior side wall serving as the side wall of two 
settings. In the case of the boiler last described the neces- 
sity for side cleaning doors makes it impossible to join more 
than two boilers in this way. 

The Heine boiler is supported by standing the front 
and rear headers upon the brickwork of the setting and 
it can therefore expand freely in all directions. 

A section of the Sterling water-tube boiler is shown in 
Fig. 233. This boiler consists of three upper horizontal 
drums connected by short curved tubes and connected to 
a single lower horizontal drum by means of long tubes which 
are curved near the ends. The curves of all tubes are so 
made that the tubes enter the drum surfaces radially, thus 
giving a simple joint which is readily made tight by expand- 
ing the tube into the sheet. 

The feed water is introduced into the upper rear drum, 
and is gradually heated and partly purified as it passes 
downward to the lower drum, in which the greater part of 
the material precipitated from the water is caught and 



STEAM BOILERS 



381 



stored until blown off. From the lower drum the water 
is supposed to pass upward through the front bank of tubes, 
the steam formed passing to the central drum through the 
upper set of short curved tubes, and the water which is 
not evaporated passing to the central drum through the 



Feed Water Inlet 



- Steam Connection 



Smoke 
Connection 



Damper 




- 1 



Bottom Blow-off 



Fig. 233.— Section of Sterling Boiler. 

lower set of curved tubes. This water passes from the upper 
central drum to the lower and returns through the front 
bank of tubes. Any steam formed in the rear bank of tubes 
or in the rear drum passes to the central drum through the 
short curved tubes connecting the steam spaces. 

The entire boiling vessel is hung from a frame of struc- 
turah steel by means of the upper drums, so that the lower 



382 STEAM POWER 

drum hangs practically free on the tubes. Independent 
expansion of all the members is insured by this method 
of suspension and by the curvature of the tubes, which per- 
mits each one of them to bend to the extent necessary to 
equalize any strains caused by changing temperatures. 

The interiors of the tubes are cleaned by means of tools 
lowered from inside the upper drums and the exterior 
surfaces are blown off by steam jets introduced through 
doors in the brickwork of the setting. 

The Wickes vertical water-tube boiler is shown in 
section in Fig. 234. It consists of an upper and lower cir- 
cular drum, connected by straight tubes expanded into the 
lower and upper heads of the drums respectively. A 
vertical baffle placed in the center of the bank of tubes 
gives an upward path to the products of combustion when 
passing over the front tubes and a downward path when 
passing over the rear tubes. 

The feed water is generally introduced at the rear of 
the upper drum, the circulation being downward in the rear 
tubes and upward in the front tubes. 

The interior surfaces of the tubes are cleaned by tools 
lowered into them by a man standing within the upper 
drum, which is made high enough to make this possible. 
The external surfaces are cleaned by steam jets inserted 
through doors in the brickwork. 

The entire boiler is supported on brackets riveted to 
the lower or mud drum and is free to expand in all directions, 
the brickwork simply enclosing but not confining it. 

153. Boiler Rating. Practically all apparatus which is 
connected with the development of power is given a horse- 
power rating. In some cases such a method of rating is 
convenient and simple, in others it is inconvenient, irra- 
tional and complicated. The term horse-power, when used 
as a measure of work or power, means very definitely the 
equivalent of 33,000 ft.-lbs. per minute. When, however, 
a certain number of horse-power is used as the rating of a 



STEAM BOILERS 



383 



team Connection 




.Smoke 

. Connection 



7 Bottom 
^Blow-off 



Fig. 234.— Wickes Vertical Boiler. 



384 STEAM POWER 

particular piece of apparatus, it generally means that that 
piece of apparatus, when working at about its best effi~ 
ciency, can do what is necessary to make available the 
stated number of horse-power in the plant of which it forms 
a part. 

Thus a boiler rated at a certain horse-power was origi- 
nally supposed to be able to supply the amount of steam 
required by an average engine developing that quantity 
of power and to do this when working at its best efficiency. 
The water rates of engines are, however, so different that 
there is no real connection between boiler horse-power and 
engine horse-power, and it is best to consider the boiler 
horse-power as a perfectly arbitrary unit defined in a certain 
way. 

The American Society of Mechanical Engineers has 
denned the boiler horse-power as the equivalent of the evapora- 
tion o/34.5 lbs. of water per hour from and at 212° F. This 
means the conversion per hour of 34.5 lbs. of water at 212° F. 
into steam at the same temperature and therefore at atmos- 
pheric pressure. 

Each pound of steam generated under these conditions 
requires the expenditure of the latent heat of vaporization 
£.t atmospheric pressure, which is equal to 970.4 B.t.u. 
according to the latest steam tables. The older tables gave 
965.7. This quantity of heat is known as a Unit of Evapora- 
tion and is abbreviated U.E. The boiler horse-power is, 
therefore, the equivalent of 34.5 U.E. per hour or 34.5 X 970.4 
= 33,479 B.t.u. per hour. 

As practically no power-plant boilers receive their feed 
water at a temperature of 212° F. and convert it into steam 
at the same temperature, it is necessary to convert the 
weight actually evaporated to what it would have been 
from and at 212° F. and then to divide this figure by 34.5 
in order to find the boiler horse-power developed. 

The number of pounds which would have been evaporated 
from and at 212° F. if the same amount of heat had been 



STEAM BOILERS 385 

transmitted is known as the equivalent evaporation, or as 

the equivalent weight of water evaporated into dry steam from 
and at 212° F. 

The method of obtaining the equivalent evaporation 
has been defined by the American Society of Mechanical 
Engineers. The heat given to each pound of dry saturated 
steam produced is to be determined; this is to be multiplied 
by the total weight of dry saturated steam generated per 
hour, and the product is to be divided by the latent heat 
of vaporization at 212° F. Thus, for a boiler receiving 
its feed water at some temperature t f above 32° F., the 
water contains a quantity of heat equal to q f B.t.u. per 
pound, qf being found in the steam table opposite the tem- 
perature t f . Each pound of dry saturated steam leaving 
the boiler carries with it an amount of heat equal to X for 
the existing temperature. The heat supplied each pound 
in the boiler must therefore be X— q f and, for W pounds 
per hour, the heat supplied would be W(\—q/). The 
equivalent evaporation is then given by 

Equiv. evsip. = Wl^^~) lbs. per hour. . (101) 

This expression may be regarded as consisting of two 
factors, the weight of dry steam generated per hour, and 
a fraction which will always have the same value for a given 
combination of pressure and feed-water temperature. This 
fraction is called the factor of evaporation, and it is cus- 
tomary to tabulate the various values of the factor of 
evaporation for different common combinations of pressure 
and feed- water temperature. 

It should be noted that the equivalent evaporation as 
defined above gives the boiler no credit for heat given to 
water which leaves the. boiler as water, nor does it give 
credit for any superheating. The former may be justified 
by saying that the boiler, as a commercial piece of apparatus, 
is not intended to supply hot water; but many commercial 



386 STEAM POWER 

boilers are expected to supply superheated steam and should 
be given credit for heat used in that way. 

Returning now to the boiler horse-power, its value can 
obviously be found for any given boiler by dividing the 
equivalent evaporation per hour by the number 34.5. 

Boilers are supposed to be so rated that they will 
develop their rated horse-power when operating at about 
their best efficiency and will do it with moderate draft 
and reasonably good firing with average fuel. Experience 
has shown that for most boilers the best efficiency is ob- 
tained when an equivalent evaporation of from 3 to 3.5 lbs. 
of water occurs per square foot of heating surface. The 
heating surface is generally taken as the total surface in 
contact with hot gases excepting in the case of tubes. The 
outer surfaces of tubes are generally counted even if they 
be in contact with the water. An equivalent evaporation 
of 3 to 3.5 lbs. per square foot would call for a heating 
surface of from 12 to 10 sq.ft. per boiler horse-power. 

Most water-tube boilers are given 10 sq.ft. of heating 
surface per rated boiler horse-power, and most return- 
tubular boilers are supplied with 11 to 12 sq.ft. Scotch 
marine boilers are generally designed on a basis of about 
8 sq.ft. per rated boiler horse-power. 

The quantity of water which can be evaporated per 
square foot seems to depend to a great extent upon the rate 
at which hot gases can be passed over the heating surface, 
and experiments have shown that from five to eight times 
the ordinary rates of evaporation can be attained if suf- 
ficient fuel can be burned. As the rate of evaporation per 
square foot is increased above the commonly accepted 
value, the efficiency decreases, but the decrease is generally 
small for a considerable increase in rate of evaporation. 
Most power-plant boilers can give from 150 to 200 per cent 
of their normal rating, and some are now installed to operate 
for long periods at about 200 per cent of what would be 
considered a normal rating. 



STEAM BOILERS 387 

154. Boiler Efficiencies. There are a great many pos- 
sible efficiencies which may be considered in connection 
with boiler tests. The two most commonly used are defined 
by the A.S.M.E., and are: 

1. Efficiency of the boiler 

_ Heat absorbed per pound of combustible burned 
Calorific value of 1 lb. of combustible 

2. Efficiency of boiler and grate 

_ Heat absorbed per pound fuel 
Calorific value of 1 lb. of fuel' 

The names used are not very well chosen, and it is better to 
call the first the efficiency based on combustible and the 
second the efficiency based on coed. The weight of com- 
bustible burned is calculated by subtracting from the coal 
fired the total weight of moisture and the total weight of 
refuse in the ash pit. 

The heat absorbed is by definition the heat absorbed 
by the dry steam made by the boiler, but it seems probable 
that this will also be modified in the near future as suggested 
in a preceding paragraph. 

It is also possible to determine the efficiency of the grate, 
of the furnace, and of the boiling vessel, and this is some- 
times done. 

The best commercial operating values for the efficiency 
of the boiler as a whole, that is, the boiler and grate on the 
basis of total fuel fired, are about 75 per cent for good 
qualities of coal and 80 per cent for oil, but such values 
are generally obtained only in well-equipped plants operat- 
ing on comparatively constant loads. Average commercial 
values generally range from 60 to 70 per cent on a yearly 
basis in well-equipped plants which are carefully operated, 
and many boiler plants are operated at an efficiency of 50 
per cent and less. 

The pounds of water evaporated per pound of coal 



388 



STEAM POWER 



fired generally ranges between 6 and 10, and the equiva- 
lent evaporation per pound of combustible burned will 
generally fall between 8 and 12 pounds. 

155. Effects of Soot and Scale. The flue gases in real 
boilers are seldom clean mixtures of the products of com- 
bustion and nitrogen, as theory would indicate. They 
always contain more or less soot and unburned hydro- 
carbons, as well as some finely powdered ash and fuel. With 
strong draught, very large particles of ash and fuel may be 
carried by the flue gases. 

These materials are partly carried up the stack by the 
gases and partly deposited on the heating surfaces of the 
boiler. Such deposits decrease the conductivity of the 
heating surfaces, and if the deposits are heavy the loss may 
be very great. The results of one investigation on the 
effect of soot are given in Table XV, the values being 
taken from an article published in the Proceedings of the 
Institute of Marine Engineers for the year 1908. These 
values are probably too high, particularly for the thicker 
deposits, but they serve to bring out the fact that a very 
appreciable loss does occur from the presence of such 
deposits. 

TABLE XV 

Effect of Soot Deposits on Boiler Heating Surfaces 



Thickness of Deposit 
in Inches. 


Loss of Conductivity 
in Per cent. 





0.0 


1 

32" 


9.5 


1 
16 


26.2 


1 


45.2 


3 
16 


69.0 



The effect of soot deposits in decreasing the efficiency 
of boilers was used for a long time as a basis for argument 
in favor of certain types of boilers in which the heating 
surfaces were so shaped and located that such deposits 



STEAM BOILERS 389 

formed to a minimum degree and against other types 
less favorably designed from this point of view. Prac- 
tically, however, the removal of such deposits by means of 
steam jets applied at regular intervals is so simple that this 
consideration need be given little weight in the selection 
of a boiler. Provision should always be made, however, 
for the easy use of the jets for cleaning purposes. 

156. Scale. Practically all water available for boiler feed 
contains various salts in solution and it often contains solid 
matter in suspension as well. This material is all deposited 
within the boiler as the water is heated and converted into 
steam. There is thus a gradual collection within the boiler 
of all the solid material brought in by the water. 

In well-designed boilers the greater part of such deposits 
is carried to a part of the boiler in which the metallic sur- 
faces are not exposed to high temperature gases, as, for 
instance, the mud drums in water-tube boilers. It can then 
be drawn off periodically in the form of a thin mud sus- 
pended in water. In practically all boilers, however, some 
of the solid material will be carried to the heating surfaces 
exposed to high temperature gases and deposited there. 
Under the action of heat, the mud-like material gradually 
changes until, in many instances, it forms a very hard, 
stone-like coating on the heating surface. This is known 
as boiler scale. 

Such deposits may cause two kinds of trouble: They 
may decrease the conductivity of the heating surfaces and 
thus decrease the efficiency of the boiler; and, because of 
their location on the water side of the metal, they may per- 
mit the hot gases to overheat that metal, thus weakening 
it. Such overheated metal often " bags " under the high 
internal pressure and may eventually give way with disas- 
trous results. The mechanical structure of the scale seems 
to be the determining factor; scales which are easily pene- 
trated by water have little effect, while those which are very 
dense and non-permeable may cause serious trouble. 



390 STEAM POWER 

Boilers should be blown down periodically to keep them 
as free as possible of scale-forming material, and they should 
be so constructed that scale which has been formed can be 
removed easily. Very efficient tools have been developed 
for removing scale from the interior and exterior surfaces 
of tubes, so that boilers using tubular heating surfaces are 
readily cleaned of scale. 

157. Scale Prevention. Much of the solid material 
carried by water is deposited when the water is heated to 
a temperature of from 150° to 200° F., so that heating feed 
water before it is admitted to the boiler is at least a partial 
preventive in most cases. 

Nearly all of the salts which are soluble in hot water 
and therefore are not deposited when the feed is heated, 
can be made to form insoluble compounds by the addition 
of comparatively cheap chemicals. By the addition of such 
chemicals in the feed-water heaters, or in other apparatus 
specially designed for that purpose, the greater part of the 
solid content of the water can be precipitated before it is 
admitted to the boiler. 

There are a great many " boiler compounds " on the 
market which are intended to be mixed with the water as 
it is fed to the boiler and are supposed to prevent the for- 
mation of scale on the heating surfaces. All they can 
possibly do is to change the chemical composition of the 
solids; they cannot prevent the deposit of these solids within 
the boiler. They are therefore, at best, only an imperfect 
remedy. 

158. Superheaters. Many boiler plants are now ar- 
ranged to supply steam superheated 25 to 200 degrees Fahr. 
It was shown in an earlier chapter that the use of super- 
heated steam greatly improves the economy of reciprocating 
engines and turbines, and there are also other advantages 
which accrue from its use. 

Superheaters are of two kinds — separately fired and 
built-in superheaters. The separately fired superheaters are 



STEAM BOILERS 391 

enclosed in a brick setting fitted with grate and furnace 
similar to that of an ordinary boiler. The built-in super- 
heaters are installed within the boiler setting so that the 
products of combustion pass over them in flowing through 
the boiler. 

In either type the steam passes through the superheater 
on its way from the boilers to the engines. In the case 
of separately fired superheaters, the temperature of the 
superheated steam is controlled by regulation of the fire 
on the grate of the superheater, but in the built-in type 
regulation in this way is practically impossible, as the fire 
under the boiler must be controlled to suit the demand for 
steam. The control of such superheaters is therefore effected 
either by locating them in such a position that the natural 
variation in the temperature of the gases reaching them 
gives an approximate regulation, or they are installed in 
a separate chamber and hot gases passed over them in such 
proportions as necessary to give the required temperature. 

The Babcock & Wilcox superheater as applied to the 
boiler of the same make is shown in Fig. 235. The steam 
collected in the dry pipe within the drum passes downward 
to the upper manifold of the superheater and from there 
it flows through the U-shaped tubes into the lower manifold. 
From the lower manifold it flows through the superheater 
stop valve to the engine or turbine. 

The superheater is so located that the hot gases pass 
over it between the first and second passes and there is no 
way of shutting off these gases. Provision, as shown in 
the illustration, is therefore made for flooding the super- 
heater during starting, or when superheated steam is not 
desired. When flooded it becomes heating surface similar 
to that of the tubes below, the steam made passing into 
the drum through the dry pipe. 

The Heine superheater as applied to a Heine boiler is 
shown in Fig. 236. It consists of a sheet-metal header or 
box into which U-shaped tubes are expanded. The steam 



392 



STEAM POWER 



Safety Valve 



Stop Valve 




Drain 
Valve 



Front 



Gases 
from Furnaces 



Fig. 235. 



-Superheated Steam from. 
Superheater 



Saturated 

Steam to 

Superheater 




Fig. 236. — Heine Superheater. 



STEAM BOILERS 393 

enters the bottom of the header and is guided by dia- 
phragms in such a way that it passes through the lower set 
of U-tubes, returns to the header, passes through the upper 
set of tubes, and then leaves the superheater at the top. 



Fig. 237.— H.R.T. Boiler and Foster Superheater. 

This apparatus is installed in a brick chamber built into 
the boiler setting and connected with the furnace by a flue 
(not shown) in the brick side wall. A damper controls the 
flow of hot gases to this chamber and the degree of superheat 
is controlled by the position of this damper. 




Headers 



Fig. 238. — Element of Foster Superheater. 

The Foster superheater is shown installed in the setting 
of an H.R.T. boiler in Fig. 237 and the details of the con- 
struction of one element are shown in Fig. 238. The core 
is used to spread the steam in a thin stream, thus bringing 
it into better contact with the beating surface. The fins 



394 STEAM POWER 

on the exterior of the element are used for the purpose of 
getting a more extended metallic surface in contact with 
the hot gases. 

159. Draft Apparatus. Attention was called in a pre- 
ceding paragraph to the fact that there must be a difference 
of pressure between the spaces below and above the fuel 
bed in order to cause the necessary air to flow through the 
bed. This difference of pressure is called the draft. 

As a matter of fact, a slight difference of pressure is 
required to cause the flow of gases through any part of the 
boiler and the drop in pressure through the fuel bed is only 
part of the total draft required. 

The draft may be created in two distinctly different 
ways. It may be caused by a chimney or stack, and is 
then known as natural draft, or it may be produced by fans 
or blowers, in which case it is called mechanical draft. 

(a) Chimneys or Stacks. Stacks are practically always 
used in small plants because of the simplicity resulting from 
their use and because the interest on the investment com- 
pares favorably with interest on investment plus cost of 
operation for mechanical draft. In large plants fitted with 
some types of mechanical stokers, or where fuel is to be 
burned at a high rate, or where the flue gases are to be used 
for heating feed water, mechanical draft is generally installed. 
A stack of some sort is necessary even though mechanical 
draft be used, because the products of combustion must be 
discharged at a sufficient elevation to prevent their being 
a public nuisance. 

A chimney serves to carry away the hot products of 
combustion and when in operation is filled with a column of 
gases with higher average temperature than that of the 
surrounding air. As a result the density of gases within 
the stack is less than the density of the outer air and the 
gas pressure at the bottom of the structure is less inside 
the stack than it is outside. If an opening is made at this 
point, the externa] air will therefore flow in. By arranging 



STEAM BOILERS 



395 



the apparatus as shown in Fig. 239, the temperature 
of the air flowing into the bottom of the stack is raised 
as it passes through the furnace and the flow is thus made 
continuous. 

The height of the chimney determines the draft created 
by it with flue gases of a given temperature, and, with any 
given height, the area determines the quantity of gas which 




Atmospheric Pressure 
less amount caused by 
low density of hot gases 
in stack. 



Fig. 239. — Diagrammatic Arrangement of Stack. 



can be carried off in a given time. The proportions of 
chimneys can be determined from rational formulas based 
on theoretical considerations, but it is necessary to assume 
values for a number of constants and a proper choice de- 
pends largely upon experience. 

As a result, all but the more important chimneys are 
generally designed on an empirical basis and many formulas 
have been developed for this purpose. One of the most 
common methods of design is to choose the height in accord- 



396 STEAM POWER 

ance with the values given in Table XVI, and then to 
determine the sectional area according to an empirical 
assumption or formula. 

TABLE XVI 

Common Heights of Chimneys 

(Applicable to plants smaller than about 700 H.P. Larger installations should 
have stacks of from 150 to 175 feet in height unless local conditions call for 
greater height.) 



Character of Fuel. 


Height above Grate in Feet. 


Free-burning bituminous 

Anthracite, medium and large sizes 

Slow-burning bituminous 

Anthracite, pea size 

Anthracite, buckwheat sizes 


80 
100 
120 
130 
150 



Thus, some designers simply assume the sectional area 
at the top of the stack equal to about one-ninth of the grate 
area for anthracite coal and equal to about one-seventh of 
the grate area for bituminous coal. Others use a formula 
developed by William Kent, which is based upon the 
assumption that the stack should be large enough to carry 
away all the gases resulting from the combustion of 5 lbs. 
of coal per rated boiler horse-power per hour. This formula 
gives the boiler horse-power which the stack can serve 
and is 

H.P.=3.33(A-0.6>/T)VS, . . . (102) 
in which 

H.P. = Rated boiler horse-power; 

A = Internal sectional area in feet of circular or square 

chimney; 
H — Height above grate in feet. 

(b) Mechanical Draft. Fans can be so used as to force 
air into the ash pit, that is, to raise the pressure on the 



STEAM BOILERS 397 

entering side of the fire. In such cases the equipment is 
said to give forced draft. Or fans may be installed at the 
discharge end of the flues and may "draw " the gases 
through the boiler by lowering the pressure within to a value 
below that of the external atmosphere. Such an instal- 
lation is said to give induced draft. 

Forced draft suffers from the disadvantage that the 
pressure within the furnace is greater than atmospheric and 
hot gases may therefore be blown out when the fire 
door is opened. On the other hand, the fan handles only 
cool air instead of hot products of combustion as in 
the case of induced draft and its useful life is therefore 
much longer. Forced draft is much more common than 
induced draft. 

Several arrangements giving balanced draft have been 
developed. With such apparatus a pressure equal to atmos- 
pheric is maintained above the fuel bed and no hot gases 
are blown out through the firing door. 

PROBLEMS 

1. The equivalent evaporation of a boiler during a certain 
test was 3450 lbs. per hour. What boiler horse-power was. de- 
veloped? 

2. A water-tube boiler with 5000 sq.ft. of heating surface 
and rated in the ordinary way gave an equivalent evaporation 
of 25,875 lbs. per hour. At what per cent of rating was the 
boiler operating? 

3. A certain boiler produced 3500 lbs. of dry steam in one 
hour from feed water at a temperature of 50° F. The steam 
pressure was 200 lbs. per square inch gauge. What was the 
equivalent evaporation? 

4. A boiler receiving water at a temperature of 250° F. con- 
verts it into superheated steam at a pressure of 210 lbs. per square 
inch gauge and a temperature of 580° F. The boiler produces 
26,000 lbs. of steam per hour. What is the equivalent evaporation 
if the boiler is given credit for all the heat given the material 
passing through it? What boiler horse-power is developed? 

5. A boiler produces 7.5 lbs. of dry steam per pound of coal 
fired. The feed- water temperature is 80° F. and the steam pres- 



398 STEAM POWER 

sure is 125 lbs. per square inch absolute. What is the equivalent 
evaporation per pound of coal? 

6. A boiler is supplied with coal which has a calorific value 
of 13,520 B.t.u. per pound. It produces 8 lbs. of dry saturated 
steam at a pressure of 150 lbs. per square inch gauge per pound 
of coal. The feed-water temperature is 70° F. What is the 
efficiency of the outfit? 



CHAPTER XVIII 
RECOVERY OF WASTE HEAT 

160. Waste Heat in Steam Plant. There are two great 

heat wastes in the steam plant — the waste in the hot gases 
going up the stack and the waste in exhaust steam. The 
magnitude of the stack loss can best be appreciated by 
determining an approximate value for assumed conditions. 
For this purpose assume the fuel to be pure carbon, the 
excess coefficient 1.5, average atmospheric temperature 60° 
F., average stack temperature 600° F., and no moisture in 
the air. The specific heat of the flue gases may be taken 
as constant and equal to 0.24. 

With an excess coefficient of 1.5, the total weight of 
flue gas per pound of carbon burned would be about 18.4 
lbs. and the heat carried up the stack figured above room 
temperature would be 

Stack loss = 18.4X0.24 (600-60). 

= 2380 B.t.u. per pound of C burned (approx.) 

With a calorific value of 14,600 B.t.u. per pound of carbon 
this loss would be equivalent to a little over 16 per cent 
of the total heat in the fuel. 

It would be more correct to use the temperature of 
the steam in the boiler instead of room temperature, because 
the lowest temperature theoretically attainable by gases 
passing through a boiler would be equal to that of the steam 
and water on the other side of the heating surface. Under 
ordinary conditions of operation, this method of figuring 
would give a theoretically avoidable stack loss equal to 
about 50 per cent of the figure obtained above. 

399 



400 STEAM POWER 

The magnitude of the exhaust loss can be similarly 
approximated. Assume for this purpose an engine receiving 
dry saturated steam at 115 lbs. absolute per square inch and 
exhausting it with a quality of 90 per cent at a pressure of 
15 lbs. absolute per square inch. 

The heat above 32° in the entering steam is 1188.8 B.t.u. 
per pound and the heat exhausted per pound is 1053.7. 
The heat in the exhaust represents therefore about 89 per 
cent of all the heat supplied when calculations are made 
above a temperature of 32° F. If a feed-water tempera- 
ture of 60° be assumed and heat quantities be figured 
above that datum the results are practically the same. 

There are always numerous pieces of auxiliary apparatus 
in steam plants such as boiler-feed pumps, circulating pumps, 
vacuum pumps, etc. These are often steam driven and are 
generally very uneconomical in the use of heat, so that they 
throw away in their exhaust steam large quantities of heat 
originally transferred from fuel to water and steam in the 
boiler. 

161. Utilization of Exhaust for Heating Buildings. It 
often happens that steam-power plants are located within 
or in the neighborhood of buildings requiring artificial heat 
during part of the year. In such cases the exhaust steam 
from main and auxiliary engines can generally be advan- 
tageously used for this purpose. Under particularly favor- 
able circumstances, the weight of steam required by the 
plant may equal approximately that required for heating, 
and the greater part of the exhaust could then be turned 
directly into the heating system. 

The engines in plants of this character may be regarded 
as reducing valves for the heating system, receiving steam 
at high pressure and reducing the pressure to the value best 
adapted to the heating system installed. If the com- 
paratively small losses arising from radiation from the 
engine, from friction and from the presence of hot water 
in the exhaust be neglected, all heat received by the engine 



RECOVERY OF WASTE HEAT 401 

and not turned into useful mechanical energy is made use 
of in the heating system. The engine may therefore be 
very uneconomical in the use of steam and still not cause 
a waste of fuel, provided always that the heating system 
can absorb all heat exhausted.- 

Since the demands of a heating system vary from day 
to day and since there is generally no demand for heat 
during several months of each year, it follows that a high 
degree of skill is necessary in choosing the character of the 
apparatus installed. A compromise is generally made 
between the cheap and uneconomical engine allowable during 
the coldest months and the more expensive and more 
efficient engine desirable when no heating is to be done. 

There are other cases of somewhat similar character. 
In many industries use can be made of exhaust steam for 
the heating of evaporating pans, dye vats, kilns and other 
apparatus. Steam plants of an uneconomical character may 
be very economical financially in connection with such 
industries if all or nearly all of the heat in the exhaust can 
be utilized industrially. 

162. Feed- water Heating. An examination of the steam 
table will show that the total heat above 32° F. per pound 
of saturated steam varies between 1180 and 1200 B.t.u. for 
such pressures as are commonly used in boilers. The aver- 
age temperature of water as it occurs on the surface of the 
earth is probably somewhere in the neighborhood of 60°, 
so that the heat above 32° per pound would roughly average 
27 B.t.u. A boiler receiving water at 60° and converting it 
into steam at any of the ordinary pressures must therefore 
supply over 1100 B.t.u. per pound of water. 

This immediately suggests a use for heat in exhaust 
steam. Steam exhausted into very low vacuums has a 
temperature only 10° to 30° higher than the assumed average 
natural feed temperature, but steam exhausted at atmos- 
pheric pressure has a temperature of 212° F. and could 
therefore impart large quantities of heat to water at 60° F. 



402 STEAM POWER 

Since the boiler must supply over 1100 B.t.u. per pound 
of steam made, raising the feed temperature about 11° or 
12° should effect a saving of about 1 per cent in fuel con- 
sumption. By raising the temperature from 60° to 212° 
there should therefore result* a saving of approximately 13 
to 14 per cent. 

Other advantages which would accrue from this pre- 
liminary heating of the feed water would be (1) the deposit, 
outside of the boiler, of a large amount of the solid matter 
carried by the water, (2) the use of fewer or smaller boilers, 
and (3) the reduction of the strains which occur in the metal 
of some designs when very cold feed water is used. 

Exhaust steam feed-water heaters are divided into two 
types, open and closed heaters. In open heaters the steam 
and feed water are brought into intimate contact in the 
form of jets, sheets and sprays within a vessel of appropriate 
size and shape. They are often called contact heaters. 
When the exhaust steam comes from reciprocating engines 
it always carries in suspension some of the oil used for 
lubricating the engine cylinders. If allowed to enter the 
heater, this oil would mix with the feed water and eventually 
reach the boilers, where it might cause serious damage by 
depositing upon heating surfaces exposed to the fire or to 
very hot gases. Such heaters are therefore always fitted 
with oil or grease extractors when used with reciprocating 
units. When receiving the exhaust from turbines, oil ex- 
tractors are not necessary, as no lubricant is used within 
the steam spaces of such units. 

Closed heaters consist of tubes or coils enclosed within 
a metal vessel. One medium passes through the tubes 
and the other over their outer surfaces. Such heaters are 
therefore often called non-contact heaters. 

As oil is a poor conductor of heat, the exhaust steam 
from reciprocating units should be passed through an oil 
extractor before entering a closed heater in order that the 
heating surfaces may be used to the best advantage. 



RECOVERY OF WASTE HEAT 403 

Exhaust steam feed-water heaters are often divided into 
primary and secondary heaters. This distinction has nothing 
to do with structure, being based entirely on position and 
temperature. Thus there may be available exhaust steam 
at a pressure below atmospheric, as from condensing main 
units, and exhaust steam at atmospheric pressure from non- 
condensing auxiliaries. The lower pressure steam could be 
used to heat the feed water in a primary heater and the 
higher pressure steam could then raise its temperature still 
further in a second or secondary heater. 

The other great waste, that in the stack gases, can also 
be partly eliminated by using some of it to heat the feed 
water. As the highest steam temperature ordinarily avail- 
able in the exhaust system is about 212° F., and as the 
products of combustion leaving the boilers generally have 
temperatures in the neighborhood of 600° to 700° F., it 
is evident that on a basis of temperature the hot gases 
have a decided advantage as a heating medium. On the 
other hand, the specific heat of the hot gases is low, while 
exhaust steam can give up all of its latent heat with no 
change in temperature, so that on a basis of heat avail- 
able for transmission to the water, the steam has the 
advantage. 

The waste heat in the flue gases is used for feed-water 
heating in devices known as economizers, preheaters, flue 
gas beaters, etc. These devices are now built in two radi- 
cally different forms. In one form the heater is mechani- 
cally separate from the boiler and is generally located in the 
flue beyond the boiler damper. In the other the heater 
is constructed as part of the boiler but is so arranged that 
water fed into it can be heated by escaping gases before 
mixing with the water in the main circulating system of 
the boiler. The separate form of heater is commonly 
known as an economizer while the one which is essentially 
part of the boiler is called a preheater, a preheater section, 
an integral eoonomizer, etc, 



404 STEAM POWER 

For many years economizers were all of the separate 
type and were made entirely of cast iron: They consisted 
of cast-iron tubes fastened in groups into cast-iron headers 
and so arranged in the flues that the tubes stood vertical. 
The flue gases passed over the external surfaces of the tubes 
and the water on its way to the boiler flowed through tubes 
and headers. The apparatus was generally arranged so 
that water entered at the end at which the gases left and left 
at the end at which the gases entered. This maintains 
the greatest available temperature difference between gas 
and water throughout the entire economizer and is known 
as a counterflow arrangement. 

Cast iron was used both because it is a cheap material 
and because it is highly resistant to corrosion. Practically 
all real fuels contain sulphur and some of this sulphur 
appears in the flue gases as sulphur dioxide. This gas 
dissolved in water forms sulphurous acid and when further 
oxidized yields sulphuric acid. If under any conditions 
the flue gases are cooled to the dew point while passing 
through the economizer a certain amount of acid or acidu- 
lated water is deposited on the external surfaces of the tubes 
and headers and even with cast-iron corrosion is fairly rapid 
if such deposits occur in large quantities. 

On the other hand, boiler-feed water often contains 
gases and other impurities which corrode steel rapidly if 
the water is heated in contact with it. Cast-iron tubes 
have been found to resist such corrosion to a much greater 
extent than any form of steel. 

The cast-iron economizer is therefore much safer against 
both internal and external corrosion than is an economizer 
built of steel. 

Economizers are generally connected into the system 
between the boiler-feed pump and the boiler so that the 
water within them is at a pressure at least slightly greater 
than that within the boiler. When boiler pressures were 
low little thought was given to this fact but as boiler pres- 



RECOVERY OF WASTE HEAT 405 

sures were increased many engineers questioned the prac- 
tice of using cast-iron economizers under full boiler pressure. 
Opinions as to the limiting permissible pressure within 
cast-iron economizers varies, being placed by different 
engineers at all values between 150 and 250 pounds. Boiler 
pressures have now reached 300 pounds, and higher in the 
larger and more modern plants. 

Three solutions have been developed. These are: 

1. The use of cast-iron economizers at a pressure lower 
than that in the boiler, a pump drawing hot water from the 
economizer and pumping it into the boiler. This involves 
the use of two sets of pumps, one set to force the water 
into the economizer at a pressure sufficiently high to prevent 
vaporization when the temperature of the water is raised 
and the other to raise the pressure from that in the econom- 
izer to the pressure of the boiler. 

2. The use of steel economizers at full pressure. The 
use n£ such equipment involves operation of such character 
as to guard against the existence of conditions leading to 
internal and external corrosion. 

3. The use of cast-iron and steel economizers in series, the 
cast-iron economizer being operated under moderate pressure 
and at the lower temperatures at which the tendency to 
corrode is generally greatest. This arrangement entails the 
use of pumps between cast-iron and steel economizers, the 
latter being operated under full boiler pressure or higher. 

At the present time the tendency seems to be toward 
the use of cast-iron economizers under full boiler pressure 
in the plants using moderate steam pressures and in the 
smaller plants while more of the high pressure plants and 
the larger stations are adopting steel economizers operating 
under full pressure. The mixed systems and the double 
pump systems are generally regarded as undesirably costly 
and complicated. 

It is generally conceded that external corrosion of steel 
economizers can be prevented by never permitting the metal 



406 STEAM POWER 

to have too low a temperature when in contact with flue 
gases. The limiting temperature varies with the quantity 
of sulphur in the fuel but exact values are not yet available. 
A metal temperature in excess of about 125° F. seems to 
be safe for bituminous fuel with a sulphur content not in 
excess of 2 per cent while a temperature of 150 to 160° F. 
seems to be necessary with other fuels having a sulphur 
content of the order of 5 per cent. 

It is also generally conceded that under most conditions 
internal corrosion can be prevented by satisfactory degasi- 
fication of the water before entrance to the economizer. 
All water dissolves air to certain definite quantities deter- 
mined by temperature and pressure relations if the oppor- 
tunity is offered and the dissolved oxygen seems to be 
an active corroding agent when aerated water is heated 
in contact with steel. The laws governing such solution 
of air in water are such that the solubility becomes zero 
when the water is at the point of vaporization. Thus if 
water under atmospheric pressure is heated to 212° F. 
under such conditions that the dissolved air can escape 
and pass off with the vapor generated, complete degasi- 
fication can be effected. Or if water under any lower pres- 
sure is brought to the temperature of vaporization corre- 
sponding to that pressure under similar conditions, com- 
plete degasification can be brought about. This phenomenon 
at atmospheric pressure can be observed by raising a vessel 
of water to the boiling point. The bubbles which form 
first are bubbles of dissolved gas and it will be noticed that 
many of them escape from solution before any appreciable 
quantity of visible vapor, i.e., steam, is formed. 

Those plants which are using steel economizers success- 
fully are making provision for adequate degassing and for 
preventing the degassed water from again dissolving gases 
before entry to the economizer. 

In designing and operating any plant which is to use 
economizers it is necessary to strike some sort of balance 



RECOVERY OF WASTE HEAT 407 

between the amount of feed-water heating which is to be 
done by the exhaust steam and the amount which is to 
be done by gases. With the cast-iron economizers it 
has been quite common practice to heat the feed with 
exhaust steam only to that temperature required to guard 
against external corrosion with steel economizers both in- 
ternal and external corrosion must be taken into account. 
The temperatures which have beeen used successfully vary 
between about 120 and 210° F. depending upon the character 
of the fuel, the character of the water, the kind of economizer, 
the arrangement of the plant and other considerations. 

Economizers heat the feed water to temperatures vary- 
ing from about 200° to 350° F. depending upon initial 
temperature, the extent of the economizers surface, initial 
gas temperature, and other variables. The temperature 
of the flue gases drops as the temperature of the water rises, 
the numerical ratio between the two depending upon relative 
quantities and specific heats. The temperature of the gases 
leaving economizers is so low under good operating condi- 
tions that induced draft of some sort is generally necessary. 
For this reason induced draft fans are practically always 
used when economizers are installed. 

If the best results are to be obtained from the use of 
economizers the heating surfaces must be kept clean. This 
applies both to the gas side and the water side of those 
surfaces. In the case of cast-iron economizers it was cus- 
tomary to furnish mechanically operated scrapers for remov- 
ing the soot deposited on the outside surface of the tubes 
but many recent installations have omitted such scrapers 
and substituted provision for cleaning these surfaces with 
steam lances or soot blowers of some sort. Internal sur- 
faces are cleaned periodically in the same way as the sur- 
faces of boilers, the scale being removed by washing if soft, 
and by mechanical chipping devices if hard. 

The great amount of heat carried away by the flue gases 
when economizers are used would lead one to assume that 



408 STEAM POWER 

economizers should always be installed. This is, however 
far from the truth. Economizers are costly and their instal- 
lation involves the provision of space, and supporting 
structure, additional flues, amplified draft apparatus, addi- 
tional piping, etc. All this represents increased investment, 
bringing certain definite capital charges. Moreover, the 
use of economizers brings in certain operating and main- 
tenance expenses not incurred if they are not installed. 
The advisability of using economizers can be determined 
only after a proper balance of increased charges against 
the money value of the fuel saving to be expected. In 
general, the higher the cost of the fuel and the more nearly 
the plant runs throughout the year at full capacity, the 
better the chances for a net saving by the installation of 
economizers, 

PROBLEMS 

1. Determine the heat lost in the chimney gases per pound 
of coal in a plant operating under the following conditions, and 
express the loss as a percentage of the heat value of the coal. 
The coal has a calorific value of 14,000 B.t.u. per pound; the 
temperature of the gases leaving the boiler is 570° F.; 20 lbs. of 
gas result from each pound of coal burned; the mean value of 
the specific heat of the gases is 0.245; and the temperature of 
the air entering the furnace is 75° F. 

2. Determine the quantity of heat which could be obtained 
from the gases of Prob. 1 by using an economizer to reduce their 
temperature to 250° F. What percentage of the heat value of 
a pound of coal does this saving represent? 

3. The boilers of a certain plant produce 100,000 pounds of 
steam per hour when the plant is operating at full load. The 
steam-driven auxiliaries consume 10% of this steam. Steam is 
generated at a pressure of 175 lbs. per square inch gauge, and 
is superheated 150° F. The main units operate condensing and 
the condensate leaves the condensers at a temperature of 75° F. 
The auxiliaries operate non-condensing and exhaust their steam 
at atmospheric pressure and with a quality of 92%. The coal 
used has a calorific value of 13,850 B.t.u. The boiler efficiency 
/Heat given water and steam\ . 

\ Heat in fuel supplied / °' 






RECOVERY OF WASTE HEAT 409 

(a) Determine the amount of coal which would have to be 
burned per hour if the steam exhausted from the auxiliaries were 
thrown away and make-up water at a temperature of 50° F. were 
used in its place. The condensate from the condensers of the 
main unit is assumed to be returned to the boiler after being 
mixed with the make-up water. 

(6) Determine the amount of coal which would have to be 
burned per hour if the auxiliary exhaust were used to heat the con- 
densate from the main units in an open heater and if the operation 
of the plant were so perfect that no make-up water had to be 
added. 



CHAPTER XIX 
BOILER-FEED PUMPS AND OTHER AUXILIARIES 

163. Boiler-feed Pumps. The pumps used for forcing 
the feed water into boilers may be of reciprocating or 
centrifugal construction and may be driven by reciprocating 
steam cylinders, by small steam turbines or by electric 
motors. 

Steam-driven pumps are very wasteful, often using over 
100 lbs. of steam per horse-power hour. It would therefore 
seem more economical to use motor-driven pumps in electric- 
power stations, as the large power units will generate electric 
power with a consumption of from 10 to 25 lbs. of steam per 
horse-power hour and the motor efficiency will generally be 
over 80 per cent. There is, however, another point which 
must be considered. The exhaust steam from small engines 
operating boiler-feed pumps can be used for heating the 
feed water as described in the last chapter, and thus the poor 
economy of these units is of little significance; practically 
all heat exhausted can be returned to the boiler in the 
boiler feed if desirable. As a result of this considera- 
tion, coupled with others of less importance, nearly all 
boiler-feed pumps and other similar auxiliaries are steam 
driven unless there are so many that there would be 
more exhaust steam than could be absorbed by the feed 
water. 

There is at present a marked tendency toward the use 
of turbine-driven, centrifugal pumps for boiler feeding, in 
place of those driven by reciprocating steam units. The 
turbine type has several advantages, the more important 
being : 

410 



BOILER-FEED PUMPS AND OTHER AUXILIARIES 411 

(1) No oil in exhaust steam, so that latter is well adapted 
to use in all forms of feed- water heaters; 

(2) Higher speed because of continuous flow of water 
and continuous rotation of mechanical parts, thus making 
possible great decrease in size for a given amount of work, 
and 

(3) Better pump characteristics for this sort of work. 

The Duplex Steam Pump. The great majority of re- 
ciprocating steam pumps used for boiler-feed purposes are 
of the duplex pattern, one design of which is shown in Figs. 



Steam Valv.e Ciest 




Fig. 240. — Duplex Steam Pump. 

240 and 241. Two steam cylinders are arranged side by 
side, their piston rods extending into similarly arranged 
water cylinders and Carrying water plungers or pistons as 
shown in Fig. 241. As there is no rotating shaft in a pump 
of this kind, the steam valves cannot be operated by eccen- 
trics as is common with steam engines. For the purpose 
of operating these valves, bell cranks, pivoted near the 
center of length of the pump, are provided. These are 
arranged so that the long arm of one bell crank engages a 
collar on the piston rod of one steam cylinder and the short 
arm operates the valve gear of the other steam cylinder. 
The motion of the valve of one cylinder is therefore derived 



412 



STEAM POWER 



from the piston motion of the other cylinder. The steam 
pistons are practically 180° out of phase, one moving out 
while the other moves in. 

Practically no expansion of the steam is obtained in the 
cylinders of pumps of this type. They operate on the 
rectangular cycle described in an earlier chapter and are 
correspondingly wasteful in their use of steam. 



Slide Valves 




Steam End 



Fig. 241. 



Water End 

-Duplex Steam Pump. 



A turbine-driven, centrifugal boiler-feed pump is shown 
in section in Fig. 242. The turbine is a multistage arrange- 
ment of the impulse type, having one Curtiss wheel at the 
high pressure end. The pump is a three-stage device, 
the first stage discharging to the suction of the second stage 
and the second stage discharging to the suction of the third 
stage. By multistaging in this way any desired boiler- 
feed pressure can be obtained with moderate rotative speed 
and diameter, 




[To face page J t 12. 




Fig. 242. — Steam Turbine-driven Centrifugal Boiler Feed Pump. 



[To face page 412.] 



BOILEE FEED PUMPS AND OTHER AUXILIARIES 413 

164. The Steam Injector. On stc\m locomotives and 
in other portable steam plants, as well as in many small 




stationary plants, a device known as a steam injector is 
used, instead of a pump, for forcing feed water into the boiler. 
A simple form of steam injector is shown semi-diagram- 
matically in Fig. 243. 



414 STEAM POWER 

Steam from the boiler flows through the steam nozzle 
and expands from boiler pressure to a very low pressure, 
thus acquiring a high velocity at the expense of the heat 
energy which it brings from the boiler. At the end of the 
nczzle it mixes with water and imparts to that water some 
of its kinetic energy, so that the mixture moves into the 
small end of the delivery tube with a high velocity. By 
the time it has reached that point, practically all the steam 
has been condensed, and, as the sectional area of the delivery 
tube increases, the velocity of the liquid decreases with a 
corresponding increase in pressure according to Bernoulli's 
theorem. In properly designed apparatus, the resultant 
pressure is great enough to force the mixture of water and 
condensed steam into the boiler against boiler pressure. 

The space at the end of the steam nozzle is maintained 
at a low temperature by the feed water flowing through 
it and the pressure of the steam is therefore very low at this 
point, being less than atmospheric in most cases. Atmos- 
pheric pressure is therefore able to force water up the suction 
pipe if the " lift " is not too great, and when once started 
such a device can therefore " raise " its own water as well 
as deliver it against pressure. 

It is interesting to note that the efficiency of this appa- 
ratus is almost 100 per cent on a heat basis. All heat not 
radiated from the apparatus is returned to the boiler in the 
mixture of condensed steam and feed water and, as the 
external surface is very small, very little heat is lost by 
radiation. 

165. Separators. Two kinds of separators are used ki 
steam plants: (a) the oil separators already referred to for 
separating oil from exhaust steam, and (6) steam separators, 
which separate water from steam. 

As it is impossible entirely to prevent radiation from 
steam pipes, it follows that condensation will occur in any 
pipe line which carries saturated steam. Water is also 
formed in the cylinders of reciprocating engines not supplied 



BOILER-FEED PUMPS AND OTHER AUXILIARIES 415 



SieyeS 




Jacket of Insulating 
Material to Decrease 
Radiation Loss. 



Water Storage 
/ Chamber 



Fig. 244. — Steam Separator 



416 STEAM POWEE 

with very highly superheated steam, and much of it is 
generally present in the exhaust of the high and intermediate 
cylinders of multiple-expansion engines. 

A small amount of water can be passed through the 
cylinder of a reciprocating engine without mechanical 
damage, but it probably causes a loss of heat by clinging 
to the walls and assisting in the heat interchanges 
which always occur. Large quantities of water are apt 
to cause mechanical damage, as water is inelastic, and if 
more of it is trapped in a cylinder end than can be con- 
tained in the clearance, something must give way when 
the piston reaches the end of its stroke. 

It is customary to separate as much as possible of the 
water of condensation before admitting steam to the 
cylinder. The separators used are built in many different 
shapes and types, but practically all depend upon two 
principles. These are: 

(1) Water is much more dense than steam, and if a 
stream of a mixture of water and steam be made to travel 
in a curve, the water will therefore collect at the outside 
of the curve, and 

(2) Water brought into violent contact with metallic 
surfaces " wets " them and has a tendency to adhere thereto. 

In steam separators the stream of mixture is therefore 
made to change its direction of flow suddenly and to impinge 
upon baffles in such a way that the greater part of the 
liquid is caught and drained off. 

One form of separator is shown in Fig. 244. The mixture 
impinges on seives in the first part of its passage through 
the separator, part of the water passing through the open- 
ings and draining to the reservoir at the bottom of the 
device. Ridges and troughs catch all water separated and 
guide it to drains leading to the reservoir so that no water 
which is once deposited is again picked up by steam. 

Another form of separator is illustrated in Fig. 245. 
The steam impinges upon the inverted V-shaped casting 



BOILER-FEED PUMPS AND OTHER AUXILIARIES 417 

and water caught on the projecting ridges drains toward 
the sides and then downward into the receiver, while the 
steam passes on as shown. 

166. Steam Traps. In the separators just described, 
there is a constant accumulation of water which must be 
drained off periodically if the entire device is not to fill up 
and become inoperative. Similarly there is a constant 
accumulation of liquid in steam jackets, in receivers of 
multi-expansion engines and in low points in steam lines. 





CO (*) 

Fig. 245. — A Steam Separator. 

167. Steam Piping. There is a great deal of piping of 
various kinds in all steam plants and the financial success 
or failure of a plant often depends upon this apparently 
insignificant item. It is beyond the limits of a book of this 
scope to consider the many different forms of piping and the 
many different ways in which apparatus may be connected. 
This is a study in itself and one of great importance. 

It should be noted, however, that all of the following 
points must be kept in view when designing and installing 
piping and that that installation which most nearly meets 
all these requirements may be regarded as the best. 

(1) The various lines should conduct the materials flow- 



418 STEAM POWER 

ing through them with the minimum loss of pressure and 
with the minimum loss (or gain) of heat. 

(2) The pipe lines should be so constructed as to mak'3 
failure of a dangerous sort, from expansion and contraction, 
water hammer and such, most unlikely if not impossible. 

(3) All connections should be so made that the careless 
manipulation of valves cannot cause an accident. 

(4) The number of flange and screw connections and the 
number of valves and fittings should be reduced to the 
minimum, as they are often sources of weakness and are 
always costly. 

(5) The entire layout should be so arranged that inter- 
ruption of service because of pipe, or valve, failure is (as 
nearly as possible) impossible. 

(6) The cost of the system should be as small as it can 
be made, consistent with the other requirements. 

It is almost unnecessary to say that all of these desirable 
ends are never attained in any plant. A compromise must 
always be made in order to bring the cost within reasonable 
limits, but most of the recent installations show a tendency 
toward better design in this part of the plant and a con- 
sideration of reliability and safety far in excess of what was 
formerly customary. 



CHAPTER XX 
PERFORMANCE OF STEAM POWER EQUIPMENT 

168. Meaning of Performance. The term performance 
is used in a general sense in engineering and refers to the 
extent to which, or the way in which, a material or piece 
of apparatus or a structure performs, that is, the extent to 
which it meets expectations or the way in which it compares 
with other similar things. The performance of steam 
power equipment of different sorts is measured in different 
ways, depending on the purpose in view. 

Thus the performance of a given engine with respect 
to a theoretical engine might be measured by comparing 
the quantity of steam used by the real engine when doing 
a certain amount of useful work, with the quantity of steam 
which would be required if the theoretical incomplete- 
expansion cycle could really be attained and utilized. Or 
in other cases the performance might merely be expressed 
in pounds of steam used per delivered or brake horse-power 
hour or per indicated horse-power for comparison with 
other engines of like or unlike characteristics. 

Or, the efficiency of a boiler might be determined by 
dividing the heat put into water and steam (i.e., useful out- 
put or result) by the heat in the fuel fired (i.e., input). 
The numerical value thus obtained would be a measure 
of the performance of the boiler with respect to the extent 
to which it utilized thermal energy. 

Or, the engineer might be interested in a complete plant 
generating electrical energy and he might desire to compare 
it with other plants doing the same thing and particularly 
to compare it with respect to its thermal efficiency. He 
might determine the overall efficiency by dividing the 

419 



420 STEAM POWER 

thermal equivalent of the electrical output by the heat 
value of the fuel supplied to produce that output, or he might 
determine the number of thermal units supplied in fuel 
for each kilowatt-hour produced. Either expression would 
serve as a measure of the performance of the plant in this 
respect. 

Or an engineer might want to compare the performance 
of one boiler room with another with respect to the total 
of all costs. For such purposes he might use the total 
cost (including fuel, water, supplies, labor and all main- 
tenance charges) per thousand pounds of steam produced. 
Such a figure would serve as a measure of performance for 
such a purpose. 

It is impossible to consider in a book of this kind all 
or even many of the different measures of performance 
used in connection with steam power equipment. Attention 
will be limited to some of the most common which all fall 
under the heading of efficiencies. 

Reduced to its simplest terms the function of a steam 
power plant consisting of steam boilers, prime movers and 
all associated apparatus, equipment and buildings is to 
produce a certain amount of useful mechanical energy as a 
result of expending a greater quantity of heat which is 
liberated by burning fuel. The extent to which it succeeds 
in a thermal sense is measured by the overall thermal effi- 
ciency of the plant, that is the quotient obtained by dividing 
the heat equivalent of the useful mechanical energy by the 
heat supplied in fuel. This thermal efficiency is not, by 
itself, a complete measure of the performance, since high 
efficiency does not necessarily mean lowest cost of power. 
This fact should never be lost sight of. 

The overall thermal efficiency of the plant is a com- 
posite made up of a number of other thermal efficiencies, 
principally those of the boilers and those of the prime 
movers. High boiler efficiency combined with high prime 
mover efficiencv does not necessarilv mean hi.2:h overall 



PERFORMANCE OF STEAM POWER EQUIPMENT 421 

efficiency of the plant because the overall efficiency of the 
plant is also influenced by the ways in which auxiliary 
equipment is arranged and operated. However, with other 
things equal, or nearly so, the thermal efficiencies of boilers 
and prime movers will determine the over-all performance. 

These facts lead to frequent determinations of such 
thermal efficiencies for the purpose of guiding manufac- 
turers of such equipment, designers and builders of power 
plants and operators of power plants. It is the purpose 
of this chapter to consider briefly how such thermal effi- 
ciencies are determined. In order to be logical the boiler 
will be considered first and the prime mover second. 

169. Determination of Boiler Performance. The thermal 
performance of the boiler is most commonly expressed as 
overall thermal efficiency. This means the ratio of heat 
put into water and steam in the boiler to the heat in the 
fuel supplied to the boiler furnace. Quite obviously some 
means must be available for determining not only the total 
quantity of fuel supplied in a given time but also the heat 
value, or calorific value, of that fuel. The quantity of 
fuel supplied is determined by weighing the fuel if it is solid 
and by means of appropriate meters if it be liquid or gas- 
eous. Weighing is also sometimes used with liquid fuels. 

The calorific value of the fuel is determined by testing 
carefully taken samples in fuel calorimeters which will be 
described in a later paragraph. 

With the heat supplied determined in this way it is still 
necessary to measure the useful output. This is deter- 
mined from (a) the quantity of steam leaving the boiler 
and (6) the heat added to it in the boiler. The quantity 
of steam produced is usually determined by measuring the 
water supplied the boiler, either by weighing or by means 
of carefully calibrated meters and then carefully guarding 
against loss through blowoff connections or other openings 
into the boiler. 

The temperature of the feed water is measured by means 



422 STEAM POWER 

of an accurate thermometer as it enters the boiler to obtain 
a measure of the heat supplied in the feed water. The 
pressure and temperature or quality of the steam leaving 
are also determined. If the boiler is producing super- 
heated steam, pressure and temperature give all necessary 
data for determining the heat above 32° F. If the boiler 
is producing saturated steam, pressure and quality give 
the necessary information. 

Temperature of superheated steam is determined by 
means of a thermometer immersed in liquid contained in 
a well projecting into the interior of the steam pipe. Quality 
of saturated steam is determined by means of steam calorim- 
eters which are described in a later paragraph. 

170. Fuel Calorimeters. The calorific value of fuels is 
determined by means of instruments known as fuel calorim- 
eters, which were briefly referred to in Chapter XVI. 
The fuel calorimeter is a device in which a known quantity 
of fuel can be completely burned under such conditions 
that all the heat liberated can be measured. 

The bomb calorimeter is the standard instrument for 
determining the calorific value of solid fuels such as coal 
and it is also sometimes used with liquid fuels. A vertical 
section through one make of bomb calorimeter is shown 
in Fig. 246. 

In this figure the letter a designates the bomb. This 
is a heavy walled vessel, generally made of steel and lined 
with a non-corrosive material such as nickel. The non- 
corrosive lining is used to prevent attack by acids formed 
during combustion of the fuel. 

Within the bomb is supported a small shallow vessel 
known as the crucible. This crucible receives the sample 
of fuel which is to be burned in the calorimeter. The sample 
is finely ground and sometimes dried before insertion in 
the crucible and an amount weighing about one gram is 
ordinarily used. 

When making a determination the sample is placed 



PERFORMANCE OF STEAM POWER EQUIPMENT 423 



in the crucible, the crucible is placed on its support in the 
lower half of the bomb, the upper half of the bomb is put 
in place on the lower half and the two halves are locked 
together so as to form a gas-tight joint. The necessary 
oxygen for supporting combustion is insured by filling the 
bomb with oxygen under a 
pressure of several hundred 
pounds. 

Combustion is started elec- 
trically and it is then self con- 
tinuing until all of the com- 
bustible is consumed. Two 
different methods are used for 
initiating combustion. In one, 
a fine iron wire partly immersed 
in the coal in the crucible is 
heated by passing current 
through it until it ignites and 
burns and thus ignites the coal. 
In the other a platinum wire 
similarly located is heated in 
the same way. It ignites the 
coal but is not itself consumed. 

When the bomb has been Fig. 246.- 
prepared for a determination 

by the insertion of the sample and the necessary oxygen 
it is placed in the vessel b, figure 246, which is then filled 
with a known weight of water. A thermometer, designated 
by c, in the figure is provided for reading the temperature 
of the water and a motor-driven stirrer designated by d, 
is arranged to circulate the water and thus maintain it at 
a uniform temperature. 

The bomb, the water around it, the vessel b, the ther- 
mometer and the stirrer really form the calorimeter and 
could be used as such. However, it is customary to supply 
another vessel with double bottom and double walls to serve 




-A Bomb Calorimeter. 



424 STEAM POWER 

as a thermal shield for the calorimeter proper. In the 
type shown in Fig. 246 this double- walled vessel is desig- 
nated by e and is filled with water. 

It is obvious that if a sample of fuel is burned in a device 
such as that shown, the heat which is liberated will heat 
the contents of the bomb, the bomb itself, the water in which 
it is immersed, the container or vessel b and parts of the 
thermometer and the stirrer. If no heat were used in any 
other way it would be possible to determine the total quan- 
tity liberated during combustion from the respective weights 
of the materials heated and the temperature rise as deter- 
mined by the thermometer c. 

In fact, however, the interchange of heat between the 
calorimeter and its surroundings and certain other phenom- 
ena make it necessary to add complications to what would 
otherwise be a simple procedure. 

The interchange of heat between the calorimeter and 
its surroundings is spoken of as " radiation " and a " radi- 
ation correction " is determined during each test. The 
determination is made by observing the rate of change of 
temperature of the water as indicated by the thermometer 
c after assembling the apparatus and while it is attempting 
to reach the same temperature as that of its surroundings. 
When the rate of change has been established by several 
readings at regular intervals the charge is ignited. Tem- 
perature readings are continued at certain definite intervals 
while the temperature of the water rises to a maximum due 
to the heat liberated within the bomb and then begins to 
drop again at a regular rate due to loss of heat to the sur- 
roundings. 

A radiation correction is calculated from the respective 
rates of change of the temperature before combustion and 
after the attainment of maximum temperature. By means 
of this correction, the temperature rise resulting from 
combustion (as read on the thermometer) is corrected 
to what it would have been had it been possible to thermally 



PERFOEMANCE OF STEAM POWER EQUIPMENT 425 

isolate the calorimeter. With this corrected temperature 
rise it is then possible to calculate the total quantity of 
heat liberated within the bomb. 

Instead of using the respective weights of all of the 
different materials of the calorimeter in this calculation 
it is customary to determine, once for all, what is known 
as the " water equivalent " of the calorimeter. The water 
equivalent is that weight of water which would experience 
the same increase in temperature as does the real calorim- 
eter with any given supply of heat. This water equiv- 
alent in pounds, multiplied by the temperature rise in 
Fahrenheit degrees (after correction for radiation) would 
then give directly the British thermal units liberated 
within the bomb. 

There are several different methods of determining the 
water equivalent. That which is generally considered the 
most satisfactory is based upon the combustion within the 
bomb of a known weight of material of known calorific 
value. When this is done, and the necessary radiation 
correction is made, part of the known amount of heat 
liberated will be accounted for by the observed temperature 
rise of the known weight of water in the calorimeter. The 
remainder is accounted for by the same temperature rise 
of the " water equivalent " of the rest of the calorimeter. 
The water equivalent can therefore be calculated with the 
data available. 

In preceding paragraphs reference has been made to 
the " heat liberated within the bomb." This wording was 
used because the heat liberated within the bomb is not 
necessarily the same as the heat resulting from combustion 
of the fuel within the bomb. For example, if an iron wire 
is used for igniting the fuel and if this iron wire burns as 
it generally does, its combustion supplies heat in addition 
to that supplied by the fuel. 

In any case, such additional heat supplies must be 
evaluated and deducted from the total absorbed by the 



426 STEAM POWER 

calorimeter in order to obtain the heat liberated by the 
combustion of the fuel. 

The bomb calorimeter, when properly constructed and 
used, is the most accurate instrument available for deter- 
mining the calorific value of solid fuels. It is, however, an 
instrument which requires very careful manipulation and 
it does not lend itself to rapid determinations. For such 
reasons many other types have been produced and some 
of them are capable of giving results which are sufficiently 
accurate for many commercial purposes. 

An entirely different type of calorimeter known as the 
Junker calorimeter is most commonly used for determin- 
ing the calorific values of gaseous fuels. It can also be 
used with many liquid fuels. Such a calorimeter arranged 
for use with gaseous fuel is shown in Fig. 247. The fuel 
gas enters the system through a very accurate meter indi- 
cated by a in the figure. From the meter it passes to the 
gasometer b which serves to maintain constant gas pressure. 
From the gasometer it passes through the tube c to a burner 
which projects into the center of the large cylinder d which 
is the calorimeter proper. 

The gas is burned within the cylinder d by means of 
air admitted through adjustable openings in the bottom 
of the cylinder. The products of combustion first pass 
upward to the top of the cylinder. They then flow down- 
ward in tubes of small diameter which pass through a space 
between the cylinder d and an inner cylindrical wall parallel 
to it. They then pass out of the instrument through the 
nozzle e which contains an adjustable damper for con- 
trolling the rate of flow. 

Water enters the instrument through the tube / being 
delivered to a small box or head tank g. An overflow 
pipe h serves to maintain a constant level in this head tank. 
From the head tank the water flows down the pipe i and 
into the calorimeter vessel through the regulating cock j. 
Within the calorimeter vessel it flows upward in the space 



PERFORMANCE OF STEAM POWER EQUIPMENT 427 

between the two concentric cylinders previously mentioned, 
entirely surrounding the small diameter pipes which carry 
the products of combustion downward. The water finally 
leaves through the discharge tube k which spills it into a 
graduated measuring cylinder I when a run is being made. 




Fig. 247. — Junker Calorimeter. 



A thermometer m gives the temperature of the water 
entering the calorimeter and another one n gives the tem- 
perature of the water leaving. 

With the quantity of water and its entering and leaving 
temperatures determined, the quantity of heat liberated 
is calculated easily. The quantity of fuel used is measured 
by the meter a and it is therefore possible to calculate the 
heat value per unit of volume or weight. 

The instrument is commonly adjusted so that the prod- 
ucts of combustion leave the nozzle e at the same temper- 



428 STEAM POWER 

ature as that with which the gas and air enter the instru- 
ment. Under such conditions the heat value determined 
is very nearly the higher heat value and it is commonly 
so regarded. 

171. Steam Calorimeters. The devices known as steam 
calorimeters are used for determining the quality of satu- 
rated steam and, occasionally, the degrees superheat of 
superheated steam. They do not measure heat as do the 
fuel calorimeters. Since they are not heat measures the 
term calorimeter is really a misnomer. They are really 
quality meters. 

There are many different varieties of steam calorimeters, 
some very crude and some very exact instruments. Only 
the three most common types will be considered here. 
These are the Separating Calorimeter, the Throttling Calorim- 
eter and the Universal Calorimeter. 

The Separating Calorimeter is, as its name implies, a 
device in which quality is determined by separating out 
the water and determining its proportion of the total. A 
section through one form of separating calorimeter is shown 
in Fig. 248. 

Steam enters the device at the top passes through a 
small separating device indicated by a. The water which is 
carried in suspension separates out at this point and gravi- 
tates to the bottom of the inner vessel or cylinder indicated 
by b in the figure. The steam which has given up its sus- 
pended moisture passes through the small openings indi- 
cated, flows through the jacket space c and out at the bottom 
of the calorimeter through a tube d which leads it beneath 
the surface of water contained in the vessel e. The opening 
through which the steam leaves the jacket of the calorim- 
eter is made of small diameter to maintain the pressure 
in the jacket at practically the same as that within the 
calorimeter. The jacket temperature is thus maintained. 

A water glass or gauge glass, /, is connected into the 
water collecting space of the calorimeter and serves to indi- 



PEEFORMANCE OF STEAM POWEE EQUIPMENT 429 

cate the height of the accumulating water. A slide, g, which 
can be moved up and down on this glass indicates the weight 
of water within the calorimeter by reference to a scale, h. 

The instrument is used by first allowing a sample of 
the steam under test to flow through it continuously and 




Fig. 248. — Separating Calorimeter. 



to run to waste until the entire instrument has been raised 
to working temperature. Water which collects in b during 
this period is drained down to a convenient point by bleeding 
through the cock i. The weight of water in the container 
e is determined by weighing or the height is noted by 
reference to the scale on the neck of the vessel. The 
height of water in the glass / is noted and then the dis- 



430 STEAM POWER 

charge tube d is dropped as quickly as possible into the 
container e. 

The instrument then continues in operation until enough 
steam has passed through to give reasonable accuracy to 
the determination. The flow of steam through the instru- 
ment is then discontinued, the tube e is withdrawn from the 
container and the quantities of water separated in the 
calorimeter and of steam condensed in the container e are 
determined. 

If the effect of radiation and other possible sources 
of error is neglected the quality of the sample is then 

W 

* = T ^— , (103) 

W+w 

in which 

x = quality of sample expressed as a decimal fraction; 
W = dry steam leaving calorimeter and condensed in 

container, measured in pounds; 
w = water collected in calorimeter, measured in 
pounds. 

It is customary to minimize heat loss from the instru- 
ment and from the pipe bringing the sample to the calorim- 
eter by covering all parts with hair felt which is an excel- 
lent thermal insulator. For exact work it is, however, 
customary to determine the heat lost by " radiation " and 
to correct for this amount. This is conveniently done by 
expressing this heat loss in terms of the amount of water 
which will collect in the calorimeter in a given time and then 
using the formula for quality in the following form : 

x =w+^> (104) 

in which 

x, W and w have the same significance as before, and 

R = weight in pounds of that part of total liquid 



PERFORMANCE OF STEAM POWER EQUIPMENT 431 

separated in calorimeter which must have 
condensed fco supply the heat lost from the 
instrument and connections. 



An improved form of separating calorimeter is shown 
in section in Fig. 249. The improvement consists in the 
use of a calibrated nozzle 
at the bottom of the jacket 
and a gauge graduated to 
read flow of material in a 
given time. When this in- 
strument is properly cali- 
brated and used it is capable 
of giving accurate results 
but the calibration should 
be checked at intervals by 
condensing the discharged 
steam. 

The separating calori- 
meter can be used satis- 
factorily to determine the 
moisture in steam over the 
entire commercial range of 
quality. However, with 
very small moisture con- 
tents it is necessary to work 
very carefully and to apply 
several corrections to ob- 
tain accurate results. For 
this reason many engineers 
prefer to use the throttling 
calorimeter under such con- 
ditions. 

A throttling calorimeter is shown in section in Fig. 250. 
Steam enters through a small nozzle a and leaves through 
a much larger opening b from which it is carried to any 




Fig. 249. — Separating Calorimeter. 



432 



STEAM POWER 



convenient point for disposal. The nozzle at a has so 
small an opening that steam cannot flow into the interior 
of the calorimeter fast enough to build up any great pres- 
sure. Under ordinary conditions of use the pressure within 
the calorimeter is equal to atmospheric or a fraction of an 




Fig. 250.— Throttling Calorimeter. 



inch of mercury above atmospheric. The steam flowing 
through the nozzle a therefore expands from the pressure in 
the pipe leading to the calorimeter to the much lower pressure 
within the calorimeter, a process analogous to that occurring 
in the nozzles of steam turbines as described in Chapter 
XIII. There is an important difference between the two 
cases which will appear in later paragraphs. 

If the calorimeter is assumed to have been in use for 
some time, so that all metal parts are heated up to working 
temperature and if loss of heat to surroundings be assumed 
to be zero, all of the energy in the steam entering through 



PERFORMANCE OF STEAM POWER EQUIPMENT 433 

the nozzle must leave with the steam issuing from the large 
opening at the bottom of the calorimeter for the simple 
reason that there is no other way in which it can escape. 

This is the principle on which the calorimeter works 
and it can be put in the form of an equation thus, energy 
entering = energy leaving. The energy in the entering 
steam consists of two parts, that determined by its pres- 
sure and quality and determinable from the steam tables 
and that due to the velocity with which it flows in the small 
pipe leading to the calorimeter. This velocity is so small 
that it can be neglected without sensible error and it is 
always so neglected. The energy entering the calorimeter 
is therefore taken as that determined by the state of the 
steam and is equal to q-\-xr per pound of wet saturated 
steam as given in equation (15) page 34. 

The energy in the steam leaving the calorimeter is 
similarly that due to the state of the steam and that due to 
its velocity of efflux from the calorimeter. This velocity 
is always small under normal conditions of use so that it 
can be neglected. The energy determined by the state 
of the steam is therefore taken as that in the steam leaving 
and as that which must be equal to the heat above 32° F. 
in the entering steam. It follows from these facts that the 
fundamental equation of the throttling calorimeter be- 
comes 

Heat above 32° F. in entering steam = heat above 32° F. 
in leaving steam if heafc loss by " radiation " is neglected. 

If the steam tables are consulted it will be discovered 
that for saturated steam the total heat above 32° F. at a 
pressure near atmospheric is always lower than that for a 
higher pressure. In fact, if the higher pressure is 50 pounds 
or more as is common in power plant work the difference 
will be comparatively great. It is therefore true that if 
there is not too much water in the high-pressure steam 
flowing to the calorimeter the low-pressure steam leaving 
it will have to be superheated to contain the necessary 



434 STEAM POWER 

heat above 32° F. This is the only condition under which 
'this form of calorimeter can be used. Under such condi- 
tions, the fundamental equation of the calorimeter becomes 

[xr+ql^ly+Cpmils-Qh, (105) 

in which subscript 1 denotes values for entering steam, sub- 
script 2 denotes values for leaving steam, / s is temperature 
of the superheated steam in the calorimeter and t v is the 
temperature of vaporization corresponding to the pressure 
in the calorimeter. From this equation, the quality x musL 
be 

[k+C pm (t s —tv)]2 — Qi 



n 



(106) 



The calorimeter is used by allowing steam to blow 
through it and noting the pressure of the sample outside 
the calorimeter, the pressure inside the calorimeter, the 
barometric pressure and the temperature of the steam 
inside the calorimeter as indicated by the thermometer 
immersed in liquid in the thermometer well projecting into 
the interior of the calorimeter. The pressure within the 
calorimeter is expressed in pounds absolute and the cor- 
responding temperature of vaporization is determined from 
the steam tables. If this is less than the temperature 
indicated by the thermometer the steam in the calorimeter 
is superheated and the quality of the entering steam may be 
determined. It is only necessary to convert the pressure 
of entering steam into pounds absolute, take the necessary 
values from the steam tables and substitute in equation 
(106) above. 

If the steam entering the calorimeter is superheated 
instead of wet there will be no indication of such a fact 
until equation (106) is solved. The value obtained for x 
will then be greater than 1. When this happens it is pos- 
sible to determine the degree of superheat, if desired. For 
this purpose take from the steam table the total heat of 



PERFORMANCE OF STEAM POWER EQUIPMENT 435 

dry saturated steam for the pressure at which the steam 
enters the condenser and subtract it from the total heat 
of the steam leaving the condenser, i.e., the right-hand 
member of eq. (105). The difference is the amount of 
heat contained in the entering steam by virtue of its super- 
heat and must be equal to [C vm (t s — t v )]\. The value of t v 
is given in the steam tables but C pm and t s are unknown. 
A solution must therefore be found by trial and error. That 
is, some value of / s is assumed, the corresponding value of 
C pm is obtained from Fig. 14 and both are substituted in 
the expression last given and the result compared with the 
amount of heat to be accounted for. If the difference is 
greater than permissible, another initial assumption is made, 
and this is continued until the desired degree of accuracy 
is obtained. 

All calculations outlined above for determining the 
quality or initial superheat by means of observations made 
with a throttling calorimeter can be eliminated by using 
the Mollier Chart, previously given as Fig. 154. It has 
been shown that values are determined merely by assuming 
the heat above 32° F. to be the same for entering and leaving 
steam and then solving for quality of entering steam after 
obtaining all other values by observation and from the 
steam tables. 

The Mollier Chart is so drawn that any change occurring 
with constant heat above 32° F. is represented by a vertical 
line. If the condition of the low pressure steam in the calorim- 
meter (and leaving the calorimeter) is spotted on this dia- 
gram it will be found to be somewhere near the middle of 
height (about 15 pounds) and to the left of the saturation 
curve. A line dropped vertically from that point to the 
curve for the initial steam pressure will show at its inter- 
section with that curve the initial condition of the steam. 

It is obvious that if the relative locations are such that 
this vertical line crosses the saturation curve the initial 
condition must have been that of wet steam, whereas if 



436 STEAM POWER 

it does not cross, the initial condition must have been that 
of superheated steam. 

The difference in the behavior of steam flowing through 
a turbine nozzle and that of steam flowing through a throt- 
tling calorimeter nozzle can now be brought out. In the 
chapter on Steam Turbines it was shown that that which 
happens in a turbine nozzle is pictured by a horizontal 
line drawn from left to right on the Mollier Chart. Inspec- 
tion of the chart will show that this means a change which 
will reduce superheat or increase wetness. And yet in the 
case of a similar nozzle in a throttling calorimeter we seem 
to meet a different condition. 

That which occurs in the nozzle is substantially the 
same in both cases. It is in what happens after the steam 
leaves the nozzle that the difference occurs. In both cases 
expansion within the nozzle is essentially isentropic and 
represented by a horizontal line drawn from left to right 
on the diagram. The steam pressure drops, the heat above 
32° F. decreases and the difference appears as kinetic energy 
of motion. In the case of a turbine nozzle the turbine 
wheel removes this kinetic or velocity energy from the steam 
and thus leaves it with the smaller heat above 32° F. with 
which it flowed out of the nozzle. 

In the case of the calorimeter the kinetic energy is con- 
verted back into heat as the jet of steam is brought to rest 
by bombarding its surroundings. The heat above 32° F. 
thus becomes the same at the end of the whole process as 
it was at the beginning merely because the kinetic energy 
developed in the nozzle is not removed as it is in the case 
of the turbine. It is obvious that a complete representation 
of what happens in the calorimeter is more complicated 
than the simple vertical line used above to obtain a solution. 
However, it is not necessary to study the intermediate 
steps of the process as initial and end conditions give all 
that is necessary for the determination of initial quality 
or superheat. 



PERFORMANCE OF STEAM POWER EQUIPMENT 437 

The throttling calorimeter has certain definite limita- 
tions. It will work with any initial superheat but if the 
initial quality is too low the steam will not be superheated 
in the calorimeter and the initial quality cannot be deter- 
mined. This can be shown by means of the Mollier Chart. 

Assume, for instance that the intitial steam pressure 
is 50 lbs. absolute, that the initial quality is 95 per cent 
and that the pressure within the calorimeter is 15 lbs. abso- 
lute. Locate the intersection of the 50 lb. and 95 per cent 
line and run vertically upward on the chart to intersection 
with the 15 lb. line. The chart indicates that the steam 
will still contain moisture when expanded into the calorim- 
eter. In fact, the chart indicates that the initial quality 
will have to be about 97.5 per cent to just dry the steam 
in the calorimeter. 

It is not safe to use a throttling calorimeter with less 
than about 10° superheat in the calorimeter so that for 
50 lbs. absolute steam pressure, the initial quality would 
have to be about 98 per cent or higher to make the throttling 
calorimeter usable. 

The percentage of water at which the throttling calorim- 
eter becomes inoperative increases with the initial pres- 
sure so that at 200 lbs. absolute a quality as low as 95 per 
cent can be determined with this instrument. 

The Universal Calorimeter is a combination of a throt- 
tling calorimeter and a separating calorimeter. It is so 
arranged that the steam passes first through the throttling 
calorimeter. If it is superheated in this part of the appa- 
ratus nothing further is required. If it is not superheated 
by the throttling calorimeter, the separating calorimeter 
removes the remaining water. The combination thus makes 
it possible to determine the quality of a sample with any 
initial quality. 

172. Determination of Engine Performance. Engines 
are generally purchased on the basis of their thermal per- 
formance, expressed as pounds of steam of a given character 



438 STEAM POWER 

which will be required per unit of output. The unit of 
output is sometimes taken as the indicated horse-power 
and sometimes as the delivered horse-power. The latter 
is the more logical since the delivered power is the useful 
output of the engine. 

The steam supplied can be determined easily when the 
engine exhausts into a surface condenser because under 
such circumstances all the steam exhausted is condensed 
to liquid water and can be weighed. When an engine is 
not fitted with a surface condenser the accurate determi- 
nation of the steam consumption is more difficult. 

The most common method is to isolate the necessary 
number of boilers and to supply all steam delivered by them 
to the engine under test. The feed water going to these 
toilers is then weighed, all losses through blowoff valves 
and other connections, safety valves, steam separators, etc., 
are deducted and the remainder is taken as the steam 
supplied the engine. 

Occasionally steam meters are used but such devices 
are not yet developed to the point where they can be regarded 
as at all comparable with tank and scales for accuracy. 

The character of steam delivered can be determined 
by pressure gage and thermometer, or steam calorimeter, 
at the engine. 

The indicated horse-power can be determined by means 
of the steam engine indicator already described in Chapter 
VIII. 

The determination of the delivered or brake horse- 
power is not always as simple a matter as might be desired. 
There are numerous ways available but many of them are 
cumbersome, or inaccurate or difficult of application, or 
have some other drawback. Some of the more common are 
considered in the following paragraphs. 

173. Determination of Delivered Horse-power. The 
devices used to determine delivered horse-power are all 
classifiable as dynamometers. There are two distinctly 



PERFORMANCE OF STEAM POWER EQUIPMENT 439 

different types known respectively as absorption dynamom- 
eters and transmission dynamometers. 

The absorption dynamometer absorbs the energy which 
it measures as is indicated by its name. The transmission 
dynamometer merely measures the energy as it passes or 
is transmitted without absorbing any of it, except a negligi- 
ble amount in its own friction. 

One of the most common forms of absorption dynamom- 
eter is a device known as the Prony brake. It is probable 
that the term brake horse-power comes from the use of such 




Fig. 251.— Prony Brake. 



a brake for determining the delivered horse-power of engines 
many years ago. One type of Prony brake as applied to a 
steam engine is shown semi-diagrammatically in Fig. 251. 

In this figure the wheel represents either the fly wheel 
of the engine, a belt wheel or the engine shaft or a special 
wheel fastened to the shaft for the purposes of the test. A 
strap a made of steel or leather or any other convenient 
material surrounds the wheel and is spaced away from its 
face at intervals by means of blocks of wood or other appro- 
priate material. A clamping device of some convenient form 
such as that shown at b is used to pull the blocks against 
the face of the wheel with any desired force. 

If the wheel is rotated and such a band without the long 
arm shown in the figure is clamped to it, the band will 



440 STEAM POWER 

rotate with the wheel. If now the band is held stationary- 
while the wheel is rotated the friction of the blocks on the 
face of the wheel will offer some resistance to the rotation 
of the wheel. In fact, power will be expended in overcoming 
this friction and the energy " consumed " will appear 
as heat. By clamping more or less tightly more or less 
power can be absorbed in this way. This part of the Prony 
brake is therefore a power absorber. 

The frictional resistance to motion is equivalent to a 
tangential force applied at the rim of the wheel in such a 
direction as to resist its motion. If the wheel moves it 
must exert a tangential force equal and opposite to this 
resisting force. The energy supplied by the wheel (and 
absorbed by the brake) is equal to the product of this force 
by the distance through which it travels. That is 

Energy in foot pounds per minute = FX2tRu, . (107) 

in which 

F = tangential force in pounds; 
R = radius of wheel in feet: 
n = revolutions per minute. 

And the power supplied by the wheel would be 



. , Fx2irRn - v 

Power mb.p. = — . . . . (108) 

1 33,000 



It thus appears that if the tangential force could be measured 
the horse-power output could be determined. The tan- 
gential force is measured indirectly by measuring the reaction 
at the knife edge c at the end of the arm which holds the band 
or brake against rotation. It is obvious that the reaction 
at the knife edge which may be called F' must be as much 
smaller than the tangential force F required to hold the 



F f 

F ~~ 


R 


F'D- 


= FR 



PERFORMANCE OF STEAM POWER EQUIPMENT 441 

brake stationary if applied at the surface of the wheel, 
as the distance D is greater than the radius R. That is 



and 



The product F'D may therefore be substituted in eq. (108) 
for FR giving 

i v. , i_ i i F'X2TrDn ,_ N 

Horse-power absorbed by brake = — , . (109) 

which is the equation for the Prony brake. 

The force or reaction F' is measured by means of scales 
as shown. It is the total reading of the scales when the 
wheel is revolving, less that part of the reading which is 
due to the weight of the brake arm and the weight of the 
pedestal on which the brake arm acts. 

The Prony brake cannot be used to absorb great quan- 
tities of energy. It can be designed to care for several 
hundred horse-power but becomes cumbersome when built 
to absorb more than about 100 horse-power. 

There are many other forms of brake and many of them 
can be used to absorb large quantities of energy. A 
brake known as the Alden brake is shown diagrammatically 
in Fig. 252. 

The disk a is keyed to the shaft and rotated by the 
engine under test. The spaces b are filled with water under 
any desired pressure and press flexible diaphragms c against 
the rotating disk. The friction between these diaphragms 
and the disk absorbs the power generated by the engine. 

The casing d to which the diaphragms are fastened is 
free to rotate and is prevented from doing so by weights 
suspended from an arm or by resting the end of the arm 



442 



STEAM POWER 



on scales as previously described. The formula is the same 
as for the Prony brake. 

A brake of this type can be made up with numerous 




Fig. 252.— Alden Brake. 



rotating disks and can thus be made to absorb large amounts 
of power. 

One of the most common forms of absorption dynamom- 
eters is a direct connected electric generator, that is, a gener- 
ator which is coupled to the shaft of the engine under test. 
If all losses in the generator are added to its electrical out- 
put the resultant sum must be the input to the generator 
and the output from the engine. The use of generators 
in this way makes possible the convenient testing of large 
engines and turbines which are used to drive direct con- 
nected generators. 

Transmission dynamometers merely pass the energy along 
and measure it while doing so instead of absorbing it as in 
the case of absorption dynamometers. Many transmission 
dynamometers have been developed but very few have 
been widely used for measuring the output of steam engines. 
One has, however, achieved some prominence in this use 
because it happens to offer about the only available means 
of determining the output of a marine engine or marine 
turbine in place in the ship and driving a propeller. 

This form is called a torsion dynamometer because its 
operation depends on the torsion or twist in a long shaft 
delivering power. Several forms are available but all 



PERFORMANCE OF STEAM POWER EQUIPMENT 443 

merely indicate or record the relative twist between points 
on the shaft surface at a known distance apart along the 
length of the shaft. The horse-power transmitted is then 
calculated by formulas involving the shaft diameter, physical 
constants of the shaft material, angle of twist, revolutions 
and length of shaft under test. 



TABLES 



446 



STEAM POWER 



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CO OrhiiOCOlM NCOMNN 
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rH OOIOOOOCO t^l>l>l>l>» 

d NHH 1-5 t-I l-H i-i rH r-i rH 


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< 


0.0000 

0.0536 
0.0913 
0.1151 
0.1313 
0.1454 

0.1559 
0.1662 
0.1747 
0.1816 
0.1883 


13 

O 

H 


g 
< 


2.1832 

2.0944 
2.0358 
2.0007 
1.9776 
1.9678 

1.9435 
1.9296 
1.9183 
1.9095 
1.9008 


to 

3 
P 

< 

§ 

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H 

W 

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2 

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rH 0001^-I>CO CO CO »0 U0 to 

O OOOiOiO OS OS OS C5 C5 


O 

H 




1073.4 

1058.3 
1047.3 
1040.0 
1035 . 1 
1030.6 

1027.2 
1023.9 
1021.1 
1018.8 
1016.5 


O) 
03 




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O OOOOSOS OSOSOS0000 

d t^i^dood cQo6coi>rH 

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o 




1073.4 

1085.4 
1094.3 
1100.1 
1104.0 
1107.5 

1110.2 
1112.8 
1115.0 
1116.7 
1118.4 


Wo 

H 


« 


Oil CiOiC*i-iG> iOHOOt)! 

CO IQt-CSO© HtNNWW 

r-i iH iH iH tH rH rH 


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Ph 
P 

CO 

to 
W 
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P4 


CO CO d 


a 


0.0886 

0.2472 

0.4893 

0.741 

0.975 

1.235 

1.467 
1.736 
1.992 
2.219 
2.467 




CO 

v L - 
-§o|t 


1 


0.1804 

0.50+ 
1.00- 
1.50+ 
2.00- 
2.50 + 

3.00- 
3.50 + 
4.00 + 
4.50 + 
5.00 + 



SATURATED STEAM TABLE 



447 



t-ie<icoTmo 


co t~ co cn>© 


i-ie*eort. 

t-I iH tH tH 


1 *" 
■ iq 


ONOOOJO 

1-1 t-i iH t-I N 


iHNCOmiO 


CO 


OscOt- CO 00 
00 "OlM COCO 


O COCO (N 

i-Hcooo 


OS t^ 
1>CM 


os go co r^ oo 

NNHCO 


CO t^ CM CO O 
t-h CO CO OS CO 


CO CO 00O CO 
CO!> 1-1 OSl> 
COHrH 


HCONWO0 

CO LO"* rH CO 


lONOOO 
CO CO CO <N 


COCO 
CM CM 


rHCO CM t-h © 
CM CM <M CO CM 


OS 00 1> coco 


NhOOtJ( 
<M CO"* i-H 00 

-* TjH 00 -* O 

CO t> CO CO CO 


i—i 00 00 O -* 

go loco cm o 
LO lo »o »o lo 


>oooco 
OS co co cm 

CO 1> COLO 
^1 T^H Tt 1 Tfl 


i> CO 

rH t-H 

"*rH 

rHrH 


CO CM rH O OS 
-* rH "* rH CO 


OOrHCONO 

oo go r^ coco 

COCO CO CO CO 










b- OS GO 00 GO 
(M -* O OS -* 

CONOrHOO 

HHNcqiM 


t-h OscocOCM 
J>1>- t^ lOCO 
rJH to cot- GO 
CM<MCMCM<M 


O CO (M 00 
OSOSOO 
<N (MCOCO 


coco 

i-H CO 

coco 


CO OS co »o lo 

CO CM 1>- i-H LO 
i-l <M (MCOCO 
CO CO CO CO CO 


CO O >0 OS CM 
OS CO CO Os CO 
co-* -* rH lo 
CO CO CO CO CO 


OOOOO 


OOOOO 


OOOO 


oo 


OOOOO 


OOOOO 


-* O GO-* CM 
LO 00 -* i— l CO 
Jh- i-h 00 CO tH 

OS OS CO CO 00 


L.8285 
L.8161 
L.8053 
L.7S68 
L.7874 


l^l> rH rH 

OS (N COO 

r^r^coco 


iCOS 
CO-* 
to to 


rH -* o oo 

OS-* O cO(M 
rH rH rH COCO 


Ot-itH OS CO 
00 tH o COCO 
CM CM CM t-h t-i 






N03 01H05 


CC*OCOH 


-* CO CM CO 


COOS 


CM LO GO O CM 


"* t^OSOCM 


i-i «* tO £~ l> 

CO CO CO CO CO 


00 O O O i-( 
CO c©l> J^J> 


iHHMN 


CM CM 


CO CO COrH -* 

r^r^l>L^J> 


tH tH -* LO LO 


OS 1>- -* CO Tf 


OrH CM-* Os 


GO oooco 


coco 


-* t-h os 00 00 


OSO CM LOGO 


CM COCO CO <M 
1> LO "* CO CO 

OS Os OS OS Oi 


i> CM CO-* O 

WINHHH 
OS OS OS OS OS 


t^rH CM OS 

OOO OS 
OS OS OS GO 


b- CO 

OS OS 
GO GO 


rH CM Osi>- lO 
OS OS CO GO GO 

oooooo oooo 


CO CO O COCO 
GO OOGOt^L^- 
GO 00 00 00 00 


CO O CO t^ CO 


00 00 CM O O 


CM CO CM OS 




CO CD1> 00 O 


MNHiOO 


1034 
1021 
1012 
1005 
1000 


OS OS GO 00 00 
OS OS OS OS Os 


OSCO-*i-H 

OS Os OS Os 


Os os 


NiOMHO 
CO CO cO cO CO 
OS OS OS OS OS 


co colo co oq 

LO LO LO LO LO 

OS OS OS Os Os 


69.8 

94.0 

109.4 

120.9 

130.1 


OSt^OOCMrH 

i> -*' O CO i i-I 

CO -* LOLOCO 

T— 1 T— 1 T-H 1— 1 1— 1 


i>OS COLO 

LOOSCO b- 
cOcOt^t^ 


oo 

o'.t-I 

00 00 


TH lO lO Tfl l— * 

•*r^ oco cd 

CO GO OS OS OS 


GOCOOOi-H-* 
00 i-H CO CO 00 

osoooo 

rH CM CM CM CM 


-* o o»o »o 


N>OOHH 


OSLOOrH 


-*1> 


O t-i CM <N Cq 


i-H O 00 CO-* 


-* iOi-h COO 
O i-h <N (N CO 


M0 05HCO 
CO COCO rH rH 


rH cOOOOs 
^r '^ ■* rp 


gg 


<M CO-* loco 
lO LO lO LO LO 


1> 00 00 OS o 

LO LO LO LO CO 


CQ1OMH00 
GOrH lOO CM 


CO iCCON <M 
O CO 00 CM CM 


to coi> lo 

l>OS00»O 


OO 

•CO 
CM i-< 


CO-* rJH <N O 


CO T-H LO GOi-H 


CO OS (M to CO 
i-H i-H CM CM CM 
CM CM (M <M CM 


O CO LO I> o 
CO CO CO CO-* 
CM CM CM <M CM 


i-H COi-HCO CM 


OCOCM 00 CO 
!>•!>. GO GO OS 


1> t-H LO OS 

OsOOO 
i-H CM <M (M 


vHNCOtJ<IO 


CO OCO O© 


tHtHtH tH 




CO t> 00 C7SO 


CJCMCMCqCM 








CO 

oo 

* 


CO CO CO CO CO 


CO CO CO CO CO 




tHCMCO"* no 


CO l> CO oso 



448 



STEAM POWER 





! 


rH^£ 


^ 


CO t- COOS O 
(NCNCNCNCO 


«th«oooo 

CO CO CO CO ^* 


WtH»00O 
t* rH tH •># IO 




m >6 


CO 


co oor-c75^ 

NhcOhN 


CO CO 00 rH OS 


(M Os O rr rH 

O tOCO 00 »0 




tC IO T* -<HH CO 


MINHHO 


oososoooo 




Pi 

o 

« 
H 


O 

S 1 


< 


(MM»ONH 
rJH 00 (M (Oh 
»OtJH t^COCO 
COCO CO CO CO 


lOl^rH lOrH 

oOHcq-* 

CN rHOOSOO 

CO COCO CO CO 


2759 
2681 
2607 
2536 
2468 










' — \ 


73 

•a 
3 


■a 

■e- 
< 


rl<^05(MO 
COOS CO to 00 
IO tO CO CO CO 
CO CO CO CO CO 


COtH coi>o 

CO 00 CO 1> CO 
t^t^ 00 O0 OS 
CO CO CO CO CO 


CO CO Ol>C0 

(OOi*NH 

OSOOOrH 
CO Tfl TtH tH tH 


1 


ooooo 


ooooo 


OOOOO 


t 

w. 
P 

13 


1 


g 
< 


CO b- 00 Cii-H 

HOOOCi 
I>1> t^t- CO 


GO rH COCO rH 
CO OSTtH OCO 
OS 00 00 00 I>- 

co CO CO CO cO 


rH CO t^ CO rH 

(NOO^rMGO 

l> CO CO CO to 
CO co CD CO CO 




00 

s 
p 

<J 
B 
« 

H 

w 
H 

W 

00 

« 

1 

H 

w 


rH 


■2-3 


* 


ICN00O5H 


COt^ OS CO "tf 


coosocoto 


ic »o tO to CO 


CO(DCONN 


l>l> oo oo oo 




s 

X 


rH tOO tOO 

t^ t^t^l> CO 

00 00 00 00 00 


CO t* oo CO OS 

coco ooo to 

CO CD CO to to 

oo oo oo oo oo 


CD CO CQ rH O 

CO rH OS l> tO 
tO tO Th TfH TH 

00 00 00 00 00 


p 


o 
H 




(ONOO^H 


tO rH 1> IO CO 


CO CO CO CO to 


H 

CQ 


O OSb- CO to 
tO tJ< tH rH H/i 

OS OS OS OS OS 


NON>0« 
tH rJH CO CO CO 
OS OS OS OS OS 


rH OS t^ tO CO 
COCO CO CO CO 
OS OS OS OS OS 




IS 


§ 

^ 


CO t^ 00 00 00 


COCO COOS rH 


rH O 00 tO rH 




OM-^ CO 00 


CO CO OS CO CO 
CO CO CO CO CO 
CO COCO CO CO 


OS CO HHt^O 
CO TjH TH -HH IO 

<M CO CO COCO 




"3 
o 


g 


CO OS COCO OS 


rH CO CO H/i TjH 


COCO O00 CD 




HHMCOM 
CO cO cO CO CO 


lOCOI>00OS 

co CD cO cO cO 


OrHCOCOCO 








- 


242.2 
244.4 
246.4 
248.4 
250.3 


HOOIMCO 

TjH 1> rH "$ 1> 

to to CO CO CO 
CO CO CO CQ CO 


CO rH OOtOO 

OCO lOOOrH 
l>t^t^t^00 
COCO CO CO CO 




e 

00 
02 
B 
« 


* oa Q 


B. 


tst-oooo 

<N CO CO <N CO 


NrHCOOOO 
00 CO CO 00 ^ 


CNItHCOOOO 




* Gauge 
Pres- 
sure. 
Lbs. 


3 
y 

2 


CO CO CO CO CO 


CO CO CO CO CO 


CO CO CO CO CO 




i-h(MC0tJH to 

rH rH rH iH i-H 


l> OS rH CO tO 
rHrHCOCOCO 


l> OS rH CO tO 

CO CO CO CO CO 



SATURATED STEAM TABLE 



449 



(N^WOOO 

to to 10 uaco 



oo t> t> t^ i> cotoocdcd cd »o 10 io »o 10 10 10 10 •*" rj< rdi rfl ^ tjh 



o 



CNOOOOOO "*Ol>COcO 00O»OO*O CCHOOH C0iO00<NI> 

OC01>i-iCO 0>00^fl5 tJHhiOh® <NOO"*OCO (NOO^HN 

^MNNH hOOOGO 0000t>-t^cO CO O »0 iO "* ■* CO CO CO <M 

INNWININ <N<N<M'-l'-l i-l r1 »-H iH t-H i-H rH r-1 rH rH rH. rH rH i-H rH 



l^O <M <M <N 

i-i i-h IM (M <N 



oco 

CO CO CO CO""*! 



NINNH lO 
CO CO 00 i-i CO 
rJH rt* "* iO >0 



t^OrH CO"* 
WNON "* 

•O to CO CO CO 



T^ ^H T^l ^^ CO 
CDGOON^t* 
CO CO t^ t^ t^- 



ooooo 


OOOOO 


ooooo 


OOOOO 


OOOOO 


6549 
6519 
6490 
6460 
6432 


fflO»OHN 
O 00 lOCO o 
"* CO COCO CO 
cO CO cO CO CO 


*OCO<M HO 
00 CO"* <M O 

CO cO cO co cO 


OOHCOiO 
00 CO-* <M O 

CO CO CO CO CO 


N05NCOO 
00 CO lOCO <M 
OOOOO 
cO cO cO co co 






COOOOrHCO 


"* COt^OOO 


0<MCO"*iO 


oooooo 


OHMM^ 


00 00 OOOO 


O O o OS o 


OOOOO 
00 00 00 00 00 


OOOOO 
00 00 00 00 00 


i— 1 rH t— 1 rH rH 

oooooooooo 


i-i i-i co ■* co 


osw iocs CO 


00(MI>CO 00 


"* OOCOO 


I> "* rH 05 CO 


co i-i 05t>- io 

"*"*COCOCO 
00 00 00 00 00 


CONOOCN 
CO CO CO <N <M 
00 00 00 00 00 


<M(M<N CSItH 

oo oooooo 00 


001> *0"* CO 

oo oooooo 00 


HOGNtO 
HH OOO 

oooooooooo 


l>OS<N »-00 


CO 00<NI><N 


00tJHO1>C0 


ON-*NO) 


r^iococ^o 


rH OOOCO"* 

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O) Qi Qi Qi Qi 


MHO00N 

rHi-lrHOO 

0)00 


*C"* CO rH © 

OOOOO 
05 Oi o o o 


UN CO 1OC0 

o O O o o 
oooo 00 00 00 


NhOOOO 
OO OOOOO 

oooooooooo 


©HIO00H 


CO "* MO CO CO 


lOOCONO 


OOiCtNQiC 


i-Hl>CO 00 CO 


<N»OI>OcM 

lO iO *C iO CO 
CM <M (M <N(N 


rJicOOOOiM 
CO CO C01>I> 
<N <N<M<N<N 


"* COOOOCM 

i>t^t>.oooo 

Cq<N(M<M<M 


CO iOi> 00 o 

oooo oo oo o 

<N <N <M(M<M 


03 CO lO CO 00 
OOOOO 
<M(M<N<N<M 


CCONiiO 


CO (M 00 CO 00 


■* os"* oo co 


00 <M COO^cH 


OOMCOOCO 




!>■ 00 00O O 


OOHHN 

oo oo oooooo 


(NCOCO"*"* 
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rH iO iO CO CO 

oooooooooo 


05INION 


OOOOO 


00I>»OCO© 


00 "*rHl> CO 


00*OtH00 



COiOGOOiM 

oo oo ooo o 

<M<M<M<N<M 



-^NftHM 
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CM <M <N COCO 



"*COOOO<N 

O O O rH i— 1 
CO CO CO CO CO 



CO»OI>. ooo 

rH rH rHrH (N 

CO COCO COCO 



rH CO"* CO!> 

<M <M <N <N <M 

CO CO CO CO CO 



04^(0 000 
00 00 00 00 o 



N rH <© 00 © 

o> o o o o 



CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO CO lO CO CO CO CO CO 



CO CO ^ ^ ^ 



lOiOOOcO «OcOI>l>l> 



450 



STEAM POWER 





Pres. 
Abs. 
Lbs. 

Sq.in. 


a 


IOOI0010 

OHHNN 


© IO© lOO 
CO CO ^ ^ to 


>o o to O IO 
tot© CO C- t» 


rail 




ONOCDM 
CO rfH 00 CO 00 
NO00N1O 


(M h O M (M 

IOC0HHH 

r« CO CO i-h © 


©^COtOCO 
NCOtONO 

05 00NOCO 


xtn tHH CO CO CO 


CO COCO CO CO 


CO CO CO CO CO 


o 
« 
& 

H 


O 

a 


< 


hoOOtHO 
OS O CO to 00 

HHOCD00 
l-H t-HtH ©© 


©co toco © 

O^NHiO 

oon ©©to 
©©©©© 


©T-H ©T-H 00 

G0CONC0 © 
"-tf rH CO CO CO 

©©©©© 






"2 

13 


"2 s 
<I 


© tH N © © 

00 CON i-l iO 

N 00 OO © © 


©con ©r« 

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■rf iO IO to iO 


tOOO ©© © 
N©CO © © 
i-H CO CO CO CO 
to to to to to 


OOOOO 


©©©©© 


©©©©© 


o 


g 
< 


oco t^co os 

00 "* O l> CO 

OS OS OS 00 00 

to to to to to 


nnn © co 

ON TjH H © 

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iO to to to to 


^ © tO©N 

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IO to to to to 






g 

s 
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W 
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CO 00 OS © t-H 


co tf ©r>oo 


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HHH(N(M 

00 00 00 00 00 


CO CO CO CO CO 

00 00 00 00 '00 


CO coco coco 

00 00 00 00 00 






©N ©CO © 


OiOONi* 


rH 00 © tO ^ 


COON»ON 
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©N lOCO © 

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Is 

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3 


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co ©© "*co 


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t^ co co © © 
00 00 00 00 00 


rH 00©TtHC0 

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a; 




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0©©i-ht-h 
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00 tH t-jh> © 

h co co coco 

CO CO CO CO CO 


CO to 00 © CO 

CO CO CO TjH rtH 
CO CO CO CO CO 


Is 
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s 


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1> 00 GO OSO 

oo oo oo oo os 


tH rH CO CO CO 

© © © © © 


tH •* to to to 

© © © © © 


ft . 


- 


TJH 00 t-H CO'* 


xH CO tH 00 to 


© © © tooo 


i-H tJH 00 tH tHH 
CO CO CO ^ ^ 
CO CO CO CO CO 


n © co to oo 

xHH IO IO tO iO 

CO CO CO CO CO 


i-H CO © 00 © 
co © © © t> 
CO CO CO CO CO 


H 

K 

tn 
« 


cd 5S 3 


a 


IOOIOOIO 

OHHNN 


©tooio© 

CO CO tJ< ^) to 


>oo too to 

to co co I> t- 




Pres- 
sure. 
Lbs. 


a 

& 
a 


CO CO CO CO CO 

© lodod 

OS OS © O t-h 


CO CO CO CO CO 

IO © iO © to 
t-h CO CO COCO . 


CO CO CO coco 
© tO©tO© 

«HH TjH LO «0 © 



SATUBATED STEAM TABLE 



451 



OlOOiOO 

COOOOOIO 



NWtHWOON 



COOO CO CD O O i-< 

COCDOrtiOS lOtONCO IC 

*0 rti rti CO <M GO iO i-h OS Tti CM O 

CM CM <N <M CM i-i r-i i-l O © C > O 



»OrH^ CO OS 

HCOHCOH 

NhhOO 
OOOOO 



OiOGOCM 

CO CM CD <M CO »0 , 

OS OS 00 00 CO -* ' 



oooooo 



00 CO'* ON 

CO COCOT^Tt* 
IO iO iO »o *0 



COOO 

l> 1> T-H 00 

CO 00 <N rJH CO - . - . 
iO »0 CO CO 1> 



OOOOO OOOOO 



CO O 00 CO CO 
^N05N>0 
IO »0 Th tJH "* 
LO »0 iO iO »0 



COOS 

J> CM OS O 

<N t-i 00 1> OS . - 

io »o -* Ttn co *~ ^ 



."^^ iocoi> co io 



TJH -*«# -* »0 O 00 



i> loeoH os 
CO CO CD CO *0 



NOHQCNOO 
^ CM ON 00 CO 
t> t^ t> CO IO Tt< 



GO 00 OS O CM CO CO O 



O OOcOiO CO 

IO Tt< TJH ^ Tt< 

GO GO GO GO GO 



COHCMWOO 
CM i-h GO CD CO OS 
00 00NNCO-* 



COO'tf NOS 

iO GO O CM ■*' 
tF -^ iO »0 to 
CO CO CO CO CO 



»OcM CM G0»O, 
COCO'* rt^ iO 



Tt*O0MNH »OH 



COCONNOO 
OS OS Os Os Os 



i-i -<* CO 00 OS T-iiOGOCO 



MIONC5H 
NNNNGO 

CO CO CO CO CO 



HN^MOOCD 
•<* ■"*'<*■<* iO CD N 



oiaoioo 
oo oo o cs o 



ooooooo 

tOOOOOOO 

NMtJUOOON 



CO CO CO CO CO 



CO 

CO CO CO CO • 

• • • • IO 

IO lOtOlO 00 

CO COOO oo os 



o 



t-1 t-1 t-1 t-1 i-i CMCMCO*** * * 



452 



STEAM POWER 



PROPERTIES OF ONE POUND OF SUPERHEATED STEAM 

[Condensed from Marks and Davis's Steam Tables and Diageams, 1909, by 
permission of the publishers, Longmans, Green & Co.] 

Sp.V. = specific volume in cu.ft.; AQ = B.t.u. total heat above 
32° F.; A<£ = total entropy above 32° F. 



Absolute 
Pressure. 
Lbs. Sq.in. 



Sat. Temp. 
°F. 



Degrees of Superheat. 







50 



100 



150 



200 



250 



300 



15 

(213) 



50 

(281) 



100 

(327.8) 



110 

(334.8) 



120 

(341.3) 



130 

(347.4) 



140 

(353.1) 



150 

(358.5) 



Sp.V. 
AQ 
A(f> 

Sp.V. 
AQ 

A<t> 

Sp.V. 
AQ 
Acj> 

Sp.V. 
AQ 
A<}> 

Sp.V. 
AQ 
A0 

Sp. V. 

AQ 
Acj> 

Sp.V. 
AQ 
A0 

Sp.V. 
AQ 
A<t> 



26.27 
1150.7 
1.7549 

8.51 
1173.6 
1 . 6581 

4.43 
1186.3 
1.6020 

4.05 
1188.0 
1 . 5942 

3.73 
1189.6 
1.5873 

3.45 
1191.0 
1.5807 

3.22 
1192.2 
1.5747 

3.01 
1193.4 
1.5692 



28.40 
1174.2 
1.7886 

9.19 

1198.8 
1.6909 

4.79 
1213 
1.6358 

4.38 
1215.9 
1 . 6282 

4.04 
1217.9 
1.6216 

3.74 
1219.7 
1.6153 

3.49 
1221.4 
1.6096 

3.27 
1223.0 
1 . 6043 



30.46 
1197.6 
1.8199 

9.84 
1223.4 
1.7211 

5.14 
1239.7 
1.6658 

4.70 
1242.0 
1 . 6583 

4.33 
1244.1 
1.6517 

4.02 
1246.1 
1.6453 

3.75 
1248.0 
1.6395 

3.51 
1249.6 
1.6343 



32.50 
1221.0 
1.8492 

10.48 
1247.7 
1.7491 

5.47 
1264.7 
1 . 6933 

5.01 
1267.1 
1.6857 

4.62 
1269.3 
1.6789 

4.28 
1271.4 
1.6724 

4.00 
1273.3 
1 . 6666 

3.75 
1275.1 
1.6612 



34.53 
1244.4 

1.8768 

11.11 
1271.8 
1.7755 

5.80 
1289.4 
1.7188 

5.31 
1291.9 
1.7110 

4.89 
1294.1 
1 . 7041 

4.54 
1296.2 
1.6976 

4.24 
1298.2 
1.6916 

3.97 
1300.0 
1 . 6862 



36.56 
1267.7 
1.9029 

11.74 
1295.8 
1.8002 

6.12 
1313.6 
1.7428 

5.61 
1316.2 
1.7350 

5.17 
1318.4 
1 . 7280 

4. 80 
1320.6 
1.7213 

4.48 
1322.6 
1.7152 

4.19 
1324.5 
1 . 7097 



38.58 
1291.1 
1.9276 

12.36 
1319.7 
1.8237 

6.44 
1337.8 
1.7656 

5.90 
1340.4 
1.7576 

5.44 
1342.7 
1.7505 

5.05 
1344.9 
1.7437 

4.71 
1346.9 
1.7376 

4.41 
1348.8 
1 . 7320 






SUPERHEATED STEAM TABLE 



453 



PROPERTIES OF ONE POUND OF SUPERHEATED STEAM 

{Continued) 



Absolute 
Pressure. 
Lbs. Sq.in. 



Sat. Temp. 

o R 



Degrees of Superheat. 



50 



100 



150 



200 



250 



300 



160 

(363.6) 



170 

(368 . 5) 



180 

(373.1) 



190 

(377.6) 



200 

(381.9) 



300 

(417.5) 



500 

(467.3) 



Sp. V. 
AQ 

A(f> 

Sp. V. 
AQ 

A<f> 

Sp. V. 
AQ 

A0 

Sp. V. 
AQ 
A0 

Sp. V. 
AQ 

A0 

Sp. V. 
AQ 
A</> 

Sp. V. 

AQ 

A(f> 



2.83 
1194.5 
1 . 5693 

2. 
1195.4 
1 . 5590 

2.53 
1196.4 
1.5543 

2.41 
1197.3 
1.5498 

2.29 
1198.1 
1.5456 

1.55 
1204.1 
1.5129 

0.93 

1210.0 

1.470 



3.07 
1224.5 
1 . 5993 

2.91 
1225.9 
1 . 5947 

2.75 
1227.2 
1.5904 

2.62 
1228.6 
1.5862 

2.49 
1229.8 
1.5823 

1.69 
1240.3 
1 . 5530 

1.03 

1256 

1.519 



3.30 
1251.3 
1 . 6292 

3.12 
1252 
1 . 6246 

2.96 
1254.3 
1.6201 

2.81 
1255.7 
1.6159 

2.68 
1257.1 
1.6120 

1.83 
1268.2 
1.5824 

1.11 

1285 

1.548 



3.53 
1276 
1.6561 

3.34 

1278 . 4 
1.6513 

3.16 
1279.9 
1.6468 

3.00 
1281.3 
1.6425 

2.86 
1282.6 
1.6385 

1.96 
1294.0 
1.6082 

1.22 

1311 

1.573 



3.74 
1301.7 
1.6810 

3.54 
1303.3 
1 . 6762 

3.35 
1304 
1.6716 

3.19 
1306.3 
1.6627 

3.04 
1307.7 
1 . 6632 

2.09 
1319.3 
1 . 6323 

1.31 

1337 

1.597 



3.95 
1326.2 
1.7043 

3.73 
1327.9 
1.6994 

3.54 
1329.5 
1.6948 

3.37 
1330.9 
1 . 6904 

3.21 
1332.4 

1 . 6862 

2.21 
1344.3 
1.6550 

1.39 

1362 

1.619 



4.15 
1350.6 
1 . 7266 

3.92 
1352.3 
1.7217 

3.72 
1353.9 
1.7169 

3.55 
1355.5 
1.7124 

3.38 
1357.0 
1.7082 

2.33 

1369.2 
1.6765 

1.47 

1388 
1.640 



INDEX 



PAGE 

Absolute pressures 41 

Absolute temperature scale 12 

Action of steam, in cylinder 24 

on impulse blades of steam turbine 237-239 

Adiabatic expansion 58 

Advance angle 167 

Advantages of condensing 261, 262 

Advantages, relative, of contact and non-contact condensers . . . 293 ,294 

Air, excess, combustion 305 

Advantages and disadvantages of 332 

Analogy, hydraulic 26 

Analyses of coal (see fuels) 320-322 

Purchase of coal on analysis 323 

Angle of advance 167 

Ash in coal 321 

Atmospheric line on indicator diagram 119 

Atoms 296 

Avogadro's Law 300 

Babcock & Wilcox superheater 391 

" water-tube boiler 375-377 

Balanced slide valves 184 

Barometer, conversion of readings from inches mercury to 

pounds per square inch 265, 266 

Barometric Condenser 271-278 

Baume scale to express gravity 303 

Bearings Ill, 112 

Bilgram diagram 168-182 

Angularity of connecting rod 179 

Diagram for both cylinder ends 177 

Exhaust and compression 175-177 

Indicator diagram from 180-183 

Piston positions ■. .'. 177-182 

Blades, impulse, action of steam on, in impulse turbine 237-239 

455 



456 INDEX 

PAGE 

Boiler-feed pumps and other auxiliaries 410-418 

Boiler, generation of steam in 38, 39 

Boilers, steam 326-398 

Circulation in 365, 366 

Classification according to — 

(1) form; (2) location of furnace; (3) use; (4) direc- 
tion of principal axis; (5) relative positions of water 

and hot gases . 326, 327 

Draft apparatus 394-397 

Chimneys or stacks 394-396 

Mechanical draft 396, 397 

Effects of soot and scale 388, 389 

Efficiencies 387, 388 

Functions of parts 327-329 

Furnaces and combustion 329-332 

Hand firing 332-336 

Mechanical grates . . 336, 337 

Mechanical stokers 338-359 

Rate of combustion 359-381 

Rating 382-386 

Boiler horse-power 384 

Equivalent evaporation 385 

Scale 389, 390 

Prevention of 390 

Smoke and its prevention 337, 338 

Strength and safety 361-365 

Superheaters — 

Built in 390, 391 

Separately fired 390, 391 

Babcock & Wilcox 391 

Foster 393, 394 

Heine 391, 392 

Types of boilers 366-382 

Babcock & ,Wilcox, water-tube 375-377 

Continental :-. 370-372 

Externally fired, return tubular 373-382 

Heine water-tube 378-380 

Internally fired, tubular 367-373 

Locomotive 370 

Scotch marine 372, 373 

Sterling water-tube 380-382 

Vertical fire-tube 367-370 

Wicks vertical water-tube. 382, 383' 



INDEX 457 



PAGE 

British Thermal unit s, 13 

Buildings, heating of, by exhaust steam 400, 401 

Built-in superheaters 390, 391 

Calorific value of coals — 

Dulong's formula 322, 323 

Fuel Calorimeter 323 

Calorific value of petroleum oils 324, 325 

Calorimeter, fuel 323, 422-428 

Calorimeter, steam 428-438 

Carbon, combustion of 298 

CO, combustion to 298-301 

C0 2 , combustion to 301, 302 

CO and C0 2 , conditions determining formation of . . . 303-305 

CO to C0 2 , combustion of 302, 303 

complete combustion of 298 

flue gases from combustion of 305, 306 

Card factors and conventional diagram 125-128 

Centigrade scale 10, 11, 12 

Chain grate stokers 339-344 

Chart— 

Mollier, for steam 230 

temperature-entropy, for steam 62-65 

Chimneys or stacks 394-396 

Circulation in boilers 365, 366 

Classification of boilers, according to — 

(1) form; (2) location of furnace; (3) use; (4) direction of 

principal axis; (5) relative position of water and hot gases . 326, 327 

Classification of steam engines 92, 93 

Clearance — steam engine — 

mechanical and volumetric 84, 85 

Clearance volume determined from diagram 131, 132 

Closed and open feed-water heaters 402-408 

Coal-fuels 318-320 

Analyses of — proximate and ultimate 320-322 

Purchase of, on analysis 323 

Coefficient, excess in combustion 305 

Combined indicator diagrams 155-158 

Combined type turbine 251 

Combustion and furnaces; steam boilers 329-332 

Combustion 296-316 

Definitions — Compounds, elements, heat or calorific value, 

atoms, molecules, etc 296-298 



458 INDEX 

PAGE 

Combustion — 
Combustion of — 

Carbon 298 

Hydrocarbons 308, 309 

Hydrogen 306-308 

Calorific value of 309 

Mixtures 309, 310 

Sulphur 309 

Combustion to — 

CO 298-301 

C0 2 301, 302 

CO to C0 2 302, 303 

Conditions determining formation of CO and C0 2 303-305 

Excess Air and excess coefficient 305 

Flue gases from combustion of carbon 305, 306 

Rate of, in boiler furnaces 359-361 

Temperature of combustion 312-314 

Theoretical temperature 313 

Commercial fuels — solid, liquid and gaseous 317, 318 

Complete expansion cycle 55-58, 72 

Complete T^-chart for steam 68-70 

Compound engine 149-151 

Compounding 141-158 

Combined indicator diagram 155-158 

Compounding 144-149 

Cylinder ratios 151-153 

Gain by expansion 141-144 

Indicator diagrams and mean pressures 153-155 

The compound engine 149-151 

Compounds — combustion 296-298 

Compression and exhaust — Bilgram diagram 175-177 

Condensation, cylinder, methods of decreasing 89-92 

Condensation, initial 81, 82 

determination of 86-89 

Condensers and related apparatus 261-295 

Advantages of condensing 261-262 

Conversion of readings from inches of mercury to 

lbs. per square inch 265, 266 

Cooling towers 294, 295 

Measurement of vacuum 262, 265 

Principle of 266-268 

Types of — 

Contact 268-278 



INDEX 459 

PAGE 

Condensers (continued) — 
Contact — 

Barometric 271-278 

Jet, Parallel flow 269-271 

Siphon 276 

Westinghouse — Leblanc 276-278 

Non-contact 278-282 

Surface 278-281 

Two-pass or double flow 280-281 

Relative advantages 293, 294 

Water required by contact condensers 291, 292 

Water required by non-contact condensers 292, 293 

Condensing, advantages of 261, 262 

Condensing plants 23 

Conditions determining formation of CO and C0 2 303-305 

Connecting rod 109, 110 

Angularity of 179 

Conservation of Energy, law of 2 

Conservation of Matter, law of 1 

Constant-quality lines on T^-chart 66, 67 

Constant volume lines, on 7V-chart 68 

Constant speed governing 218, 219 

Contact condensers 268-278 

Continental type boiler 370-372 

Conventional diagram and card factors 125-128 

Conversion of barometric readings, from inches mercury to 

pounds per square inch 265, 266 

Cooling towers 294, 295 

Corliss and other high-efficiency engines 196-215 

Locomobile type 210-215 

Non-detaching Corliss gears 201-205 

Poppet valves 205-208 

Trip-cut off Corliss 196-201 

Uniflow engine 208-212 

Corliss engine, trip-cut-off 196-201 

Corliss gears, non-detaching 201-205 

Crank end of engines 98 

Cross-head and guides 107, 108 

Cushion steam and cylinder feed 85, 86 

Cut-off governing , 215 

Cut-off ratio 128 

Cycle, area on T<£-chart representative of work 73 • 

^_ Complete expansion 55-58, 72 



460 INDEX 

PAGE 

Cycle, incomplete expansion 58-60, 74, 75 

Modifications for wet and superheated steam 73, 74 

Of events in simple steam power plant 22 

Theoretical, of steam turbine 228-231 

Cycles, desirability of various, in engines 55 

Cylinder, action of steam in 24 

Condensation, methods of decreasing 89-92 

Efficiency 139 

Feed and cushion steam 85, 86 

Ratios . 151-153 

Cylinder and steam chest 101, 102 

Decreasing cylinder condensation 89-92 

De Laval impulse turbine 239-241 

Delivered H.p., determination of 438-443 

Density, specific, of dry saturated steam 38 

Description and method of operation of D-slide valve 159-165 

Design of nozale, steam turbine 231-237 

Determination of clearance volume from diagram 131, 132 

Determination of I.h.p 120-124 

Determination of delivered H.p 438-443 

Developed horse-power 137 

Developed thermal efficiency 138 

D-slide valve 159-195 

Angle of advance 167 

Angularity of connecting-rod 179 

Bilgram diagram 169 

Description and method of operation 159-165 

Diagram for both cylinder ends 177 

Exhaust and compression 175 

Exhaust lap , 168 

Indicator diagram from Bilgram diagram 180 

Lead 166, 167 

Limitations of D-slide valve 183-185 

Piston positions 177 

Reversing engines 185-187 

Steam lap — outside lap 165, 166 

Valve setting 187-195 

D-slide valve engine, simple ■ 96-98 

Diagram, Bilgram, for both cylinder ends 177 

Bilgram, indicator diagram from 180-183 

Indicator 24 

Indicator and mean pressures for compound engines. 153-155 



INDEX 461 

PAGE 

Diagram (continued) — 
Indicator — 

combined 155-158 

Indicator, conventional and card factors 125-128 

water rate 86, 132-136 

Diagrams from real engine 192, 193 

Double acting engines 55 

Double-flow condenser 270 

Downdraft furnace 315 

Draft apparatus 394-397 

Chimneys or stacks 394-396 

Mechanical draft 396, 397 

Dry-air pump 274 

Dry-saturated steam, total heat of 33 

Specific density of 38 

Specific volume of 36-38 

Dry-vacuum pump 274 

Dulong's formula— combustion 322, 323 

Duplex steam pump 411, 412 

Eccentric 160-165 

Effective pressure, mean, methods of varying 215 

Efficiency 52, 53 

Cylinder 139 

Developed thermal 138 

Effect of temperature range on 75 

Indicated thermal 138 

Mechanical and thermal 137-140 

Of boilers 387, 388 

Relative 139 

Elements — combustion 296 

Energy- 
Conservation of energy, law of 2 

Heat 2 

Mechanical 2 

Units of 3 

Engine — 

Application of theory for an ideal to a real 54, 55 

Compound, triple, quadruple, quintuple 148 

Receiver type 149 

Tandem and cross-compound 151 

Woolf type 149 

Desirability of various cycles 55 



462 INDEX 

PAGE 

Engine (continued) — 

Double acting 55 

Efficiency 52, 53 

Heat quantities involved 50-52 

Ideal steam 43-60 

Operation of 42, 45 

Operation of the real steam 77-80 

Reversing 185-187 

Steam — 

Classification — 

(1) On basis of rotative speed; (2) Ratio of 
stroke to diameter; (3) Valve gear; (4) 
Position of longitudinal axis; (5) Num- 
ber of cylinders; (6) Cylinder arrange- 
ment; (7) Use 92, 93 

Clearance, volumetric and mechanical 84, 85 

Crosshead and guides 107, 108 

Cushion steam and cylinder feed 85, 86 

Diagram water rate 86 

Cylinder and steam chest 101, 102 

Determination of initial condensation 86-89 

Initial condensation 81 

Losses, in real installations 80-84 

Methods of decreasing cylinder condensation 89-92 

Nomenclature 98 

Principal parts 98-114 

Bearings Ill, 112 

Connecting rod 109, 110 

Crosshead and guides 107, 108 

Cylinder and steam chest 101, 102 

Flywheels 112, 113 

Frame 99, 100 

Piston 102-106 

Piston rod and tail rod 106, 107 

Shaft 110,111 

Re-evaporation in 83, 84 

Rotative and piston speed 93-96 

Simple D-slide valve 96-98 

Throttling or wire-drawing . 82, 83 

Work done by 46-60 

Engines, Corliss and other high efficiency 185-187 

Locomobile type 213-215 

Non-detaching Corliss gears. 201-205 



INDEX 463 

PAGE 

Engines (continued) — 

Poppet valves 205-208 

trip-cut-off Corliss 196-201 

Uniflow 208-212 

Entropy diagram 61-71 

of liquid, vaporization, and dry saturated steam .... 61-63 

7Vchart for steam 62-65 

Complete 68-70 

Constant quality lines 66, 67 

Diagram for a real engine 136 

Heat from 68 

Quality from 65-68 

Saturation curve 63 

Superheating lines 63, 64 

Volume from 68 

Water line . 63 

Entropy, diagrams of steam cycles 72-76 

Equivalent evaporation, boilers 361 

Excess air — combustion 286 

Advantages and disadvantages 311 

Excess coefficient 286 

Exhaust and compression — Bilgram diagram 175-177 

Exhaust lap 166, 168 

Exhaust steam, utilization of, for heating buildings 400, 401 

Expansion, adiabatic 58 

Cycle, the complete 55-58. 72 

the incomplete 58-60 

Gain by, in compounding 151 -144 

Ratio of, apparent and real 128-130 

External latent heat of vaporization 31 

Externally fired, return tubular boiler 373-382 

Fahrenheit scale 11, 12 

Feed-water heating 401 

Open and closed heaters 401-408 

Firing boilers by hand 332-336 

Fixed carbon in coal 321 

Flue gases from combustion of carbon 305, 306 

Flywheel 97, 98, 112, 113 

Regulation 216, 217 

Foot-pound, definition 3 

Forward stroke of engines 98 

Foster superheater 393, 394 



464 INDEX 

PAGE 

Frames of engines 99, 100 

Front end of engines 98 

Fuel calorimeter 323, 422-428 

Fuels. 317-325 

Commercial — 

Solid, liquid, gaseous 317, 318 

Coal 318-320 

Analyses 320-322 

Calorific value of — 

Dulong's formula 322, 323 

Fuel Calorimeter 323 

Petroleum 323-325 

Baume scale to express gravity 324 

Calorific values 324, 325 

Purchase of coal on analysis 323 

Functions of boiler parts 327-329 

Furnaces — and combustion 329-332 

— Updraft and downdraf t 336 

Gases, and vapors, steam 27 

— Flue, from combustion of carbon 305, 306 

Gaseous fuels 317, 318 

Gauge pressure 39-41 

Gearing and staging — turbines 241-248 

Gears, Corliss, non-detaching 201-205 

Generation of steam in real steam boiler 38, 39 

Generation of steam or water vapor 28 

Governing — throttle and cut-off 218 

Coefficient of regulation 219 

Constant speed 218, 219 

Governor 97, 98 

Regulation 217, 218 

Governors — 

Pendulum. 220 

Rites inertia 221-223 

Shaft 220, 221 

Grates, mechanical 336, 337 

Gridiron valve '. 185 

Guides and crosshead 107, 108 

Hand firing — steam boilers 332-336 

Heat 9 

Absorption, reversal of process. 38 



INDEX 465 

PAGE 

Heat (continued) — 

Energy 2 

Unit of 13 

From T^-chart 68 

Latent, of vaporization 30, 32 

Internal and external 30, 31 

Of liquid, q or h 31, 32 

Of superheat 34, 35 

Quantities in rectangular cycle 50-52 

Quantity of 16 

Specific 14 

Total, of dry saturated steam 33 

Of superheated steam 36 

Of wet steam 33, 34 

Value of elements and compounds 296-298 

Heat, waste — in steam plant 399, 400 

Feed-water heating 401 

Open and closed heaters 402-408 

Utilization of exhaust for heating buildings 400, 401 

Heaters, feedwater, open and closed 401-408 

Heine superheater 391, 392 

Heine water-tube boiler 378-380 

Horizontal, return, tubular boiler 373-376 

Horse-power 17 

Delivered, determination of 438-443 

Developed 137 

Hour, definition 18 

Of steam boilers 384 

Hydraulic analogy 26 

Hydrocarbons, combustion of 308, 309 

Calorific value of 309 

Hydrogen, combustion of 306-308 

I.h.p. — determination of 120-124 

Impulse steam turbine 224-228 

De Laval type 239-241 

Inclined stokers 345-348 

Incomplete expansion cycle 58-60, 74, 75 

Indicated thermal efficiency 138 

Indicator 115 

Indicator diagram 24, 115-140 

Atmospheric line 119 

Conventional and card factors 125-129 



466 INDEX 

PAGE 

Indicator diagram (continued) — 

Cut-off ratio 128 

Determination of clearance volume from diagram. 131, 132 

Determination of I.h.p 120-124 

Diagram factor or card factor 126-129 

From Bilgram diagram 180-183 

Mean effective pressure 122 

Planimeter 123 

Ratio of expansion 128-131 

Reducing mechanism 118 

Scale of spring 118 

Indicator diagrams and mean pressures for compound engines . 153-155 

Combined 155-158 

Indicator diagrams from real engine 192, 193 

Inertia governor, Rites 221-223 

Initial condensation 81, 82 

Determination of 86-89 

Injector, steam 413, 414 

Inside lap, negative 168 

Internal latent heat of vaporization 30 

Internally fired, tubular boilers 367-373 

Jet condensers 269-278 

Joule's equivalent 14 

Joule, the 3 

Kinetic mechanical energy 8 

Lap angle 166 

Lap, steam 165, 166 

Negative inside 168 

Outside and exhaust 166, 168 

Latent heat of vaporization 30, 32 

Internal and external 30, 31 

Lead 166, 167 

Leblanc — Westinghouse condenser 276-278 

Liquid fuels 317, 318 

Liquid, heat of , q or h 31, 32 

Entropy of 61 

Limitations of D-slide valve 183-185 

Balanced slide valves . 184 

Gridiron valve 185 

Piston valve 184 

Riding cut-off valves 185 



INDEX 467 

PAGE 

Locomobile type of high efficiency engines 213-215 

Locomotive type boiler 370 

Low-pressure turbines 257-258 

Matter 1 

Law of conservation of matter 1 

Units of matter 3 

Mean effective pressure 122 

Methods of varying 218 

Mean pressures and indicator diagrams for compound engines . 153-155 

Meaning of performance 419-422 

Measurement of temperature 10 

Measurement of vacuums 262-255 

Mechanical and thermal efficiencies 137-140 

Mechanical clearance, steam engine 84, 85 

Mechanical draft 396, 397 

Mechanical energy 2, 3, 7 

Potential and kinetic 7, 8 

Mechanical grates. 336, 337 

Mechanical stokers 338-359 

Mercury readings, conversion to pounds per square inch 265, 266 

Mercury thermometers 10-12 

Method of operation and description of D-slide valve 159-165 

Mixtures, combustion of 309, 310 

Moisture in coal 320 

Molecular activity 9 

Molecules 297 

Natural draft, chimneys 395 

Negative inside lap 168 

Non-condensing plants 23 

Non-contact condensers 278-282 

Surface (Wheeler) 278-281 

Non-detaching Corliss gears • 201-205 

Nozzle design, steam turbine 231-237 

Oil firing 358, 359 

Open and closed feed-water heaters 402-408 

Operation of simplified steam engine 45, 46 

Operation of real steam engine 77-80 

Outside steam lap 166 

Outstroke of engine 98 



468 INDEX 

PAGE 

Parallel-flow condenser 269-271 

Parson's type turbine 251 

Pendulum governors .....' 220 

Performance, meaning of 419-422 

Petroleum 323-325 

Baume scale to express gravity of 324 

Calorific values 324, 325 

Piping, steam 417, 418 

Piston, engine 102-106 

Piston positions for Bilgram diagram 177-182 

Piston rod and tail rod 106, 107 

Piston speeds of steam engines 93-96 

Piston valve 184 

Planimeter 123 

Plant, steam power 20 

Plants, condensing, non-condensing 23 

Poppet valves 205-208 

Positions of piston for Bilgram diagram 177-182 

Potential mechanical energy 7 

Powdered coal stokers 354-359 

Power and work 17 

Power, unit of, horse power 17 

Pressure, absolute , . 41 

Gauge. . 39-41 

Mean effective 122 

Methods of varying 215 

Pressures, mean, and indicator diagrams for compound engines. 153-155 

Prevention of smoke 337, 338 

Prime-mover 20 

Principal parts of engines 98-114 

Principal of condenser 266-268 

Properties of steam 27 

Proximate analysis of coal 320 

Pump, dry air or dry vacuum 274 

Vacuum " 282-288 

Pumps, boiler feed 410-412 

Purchase of coal on analysis 323 

Quality from TV-chart 65-68 

Constant, lines 66, 67 

Quantity of heat 16 

Rate, diagram water 86, 132-136 

Rate of combustion in boiler furnaces 3£.9-361 



INDEX 469 

PAGE 

Rating of steam boiler 382-386 

Ratio, cut-off 128 

Ratio of expansion — apparent and real 128-130 

Ratios, cylinder 151-153 

Reaction type turbine 248-251 

Receiver engine 149 

Recovery of waste heat 399-409 

Reducing mechanism 118 

Re-evaporation 83 

Regulation 216-223 

Coefficient of governor 219 

Constant speed governing 218, 219 

Governors — 

Pendulum 220. 

Rites inertia 221-223 

Shaft 220, 221 

Kinds — flywheel and governor 216-218 

Methods of varying mean effective pressure — Throt- 
tling and cut-off 218 

Relative advantages of contact and non-contact condensers.. . 293, 294 

Relative efficiency 139 

Return tubular boilers, horizontal 326, 327, 373-382 

Reversal of process of heat absorption 38 

Reversing engines 185-187 

Riding cut-off valve 185 

Rites inertia governor 221-223 

Rotative speeds of steam engines 93-96 

Safety and strength of boilers 361-365 

Saturated steam ; dry, specific volume of 36-38 

Saturated vapor 31 

Saturation curve, temperature entropy chart fcr steam 63 

for compound engine cards 156 

Scale 389, 390 

Prevention of 390 

Scale of spring, indicator 118 

Scotch marine type boiler 372, 373 

Separately fired superheaters 390, 391 

Separators 413-417 

Setting, valve 187-195 

Shaft governors 220, 221 

Shaft of engine 110, 111 

Simple D-slide valve engine 96-98 



470 INDEX 



PAGE 

Siphon condensers 266 

Slide valves 184 

Balanced 184 

Gridiron valve 185 

Piston valve 184 

Riding cut-off valve 185 

Smoke and its prevention 337, 338 

Solid fuels 317, 318 

Soot and scale, effects of, in boilers 389, 390 

Specific density of dry, saturated steam 38 

Specific heat 14 

Specific volume of dry saturated steam 36-38 

Speeds, rotative and piston, of steam engines 93-96 

Spring, scale of, indicator 118 

Stacks or chimneys 394-396 

Staging and gearing, steam turbines 241-248 

Steam, action in cylinder „ 24 

Action of, on impulse blades of turbine 237-239 

Boiler, generation of steam in 38, 39 

Calorimeter 428-438 

Consumption of steam turbines 252-257 

Cushion, and cylinder feed 85, 86 

Diagram water rate 86 

Cycles, 7>-diagrams of, 72-76 

Steam engine the ideal 43-60 

Bearings Ill, 112 

Classification 92, 93 

Connecting rod 109, 110 

Crosshead and guides 107, 108 

Cylinder and steam chest 101, 102 

Determination of initial condensation 86-89 

Flywheel and governor 97, 98 

Flywheels 112, 113 

Frame 99, 100 

Losses in real installations 80-84 

The real 77-114 

Initial condensation 81 

Re-evaporation 83, 84 

Throttling 82 

Wire-drawing 82, 83 

Methods of decreasing cylinder condensation . . . 89-92 

Nomenclature of 98 

Operation of 77-SO 



INDEX 471 

PAGE 

Steam engine (continued) — 

Piston 102-106 

Piston rod and tail rod 106, 107 

Principal parts 98-114 

Rotative and piston speeds 93-96 

Simple D-slide valve 96-98 

Steam, entropy of dry saturated 61, 62 

Generation of 28-39 

Heat of superheat 34-35 

lap, D-slide valve 165, 166 

Modification of TV>-chart for wet and superheated .... 73, 74 

Properties of 27 

Specific density of dry saturated 38 

Specific volume of dry saturated 36-38 

Temperature-entropy chart for 62-71 

7>-chart complete 68-70 

Total heat of dry saturated 33 

Total heat of wet 33, 34 

Vapors and gases 27 

Wet, effect of 53 

Steam injector 413, 414 

Steam piping 417, 418 

Steam power plant 20-22 

Steam trap 417 

Steam turbine (see Turbine) . 

Steam turbo-generators 258-260 

Stephenson link gear 186 

Sterling water-tube boiler 380-382 

Stokers, mechanical 338-359 

Chain grate 339 

Inclined, overfeed 345-348 

Powdered coal 354 

Sprinkler 339 

Underfeed 348-354 

Strength and safety of boilers 361-365 

Sulphur, combustion of 309 

Sulphur in coal 301 

Superheat, heat of 34, 35 

total heat of 36 

Superheaters — 

Built in 390, 391 

Separately fired 390, 391 

BabcQck and Wilcox 391 



472 INDEX 

PAGE 

Superheaters (continued) — 

Foster 393, 394 

Heine 391, 292 

Superheating 31 

Lines, on temperature-entropy chart for steam. . 63, 64 

Surface condensers 278-282 

Tail rod and piston rod of engine 106, 107 

Temperature 9 

Measurement of 10 

Pressure relations 29 

Temperature of combustion 312-314 

Temperature rise 314 

Theoretical 313 

Temperature-entropy chart for steam 62-71 

Complete chart 68-70 

Heat from 68 

Quality from 65-68 

Volume from 68 

T^-diagram for a real engine 136 

T^-diagrams of steam cycles 72-76 

Complete expansion cycle 72 

Area of cycle representative of work 73 

Modifications for wet and superheated steam . 73 

Temperature range, effect on efficiency 75 

Temperatures of vaporization 29 

Theoretical cycle of steam turbine 228-231 

Thermal and mechanical efficiency 137-140 

Developed thermal efficiency 138 

Indicated thermal efficiency 138 

Thermometers, mercury 10-12 

Throttle governing 215 

Throttling or wire-drawing 82 

Towers, cooling 294, 295 

Traps, steam 417 

Trip-cut-off Corliss engine 196-201 

Triple expansion 148 

Tubular boiler, horizontal return 326, 327 

Turbine, steam , 224-260 

Action of steam on impulse blades 237-239 

Combined type 251 

De Laval impulse type 239-241 

Gearing and staging 241-248 



INDEX 473 

PAGE 

Turbine (continued) — 

Impulse 224-228 

Low-pressure turbine 257-258 

Nozzle design 231-237 

Parson's type 251 

Reaction type 248-251 

Steam consumption 252-257 

Steam turbo-generators 258-260 

Theoretical cycle 228-231 

Types of boilers 366-382 

Babcock & Wilcox, water-tube 375-377 

Continental 370-372 

Externally fired, return tubular 373-382 

Heine water-tube 378-380 

Internally fired, tubular 367-373 

Locomotive 370 

Scotch marine 372, 373 

Sterling water-tube 380-382 

Wicks vertical water-tube 382, 383 

Types of condensers — 

Contact 268-278 

Barometric 271-278 

Jet, parallel flow type 269-271 

Siphon .' 276 

Westinghouse-Leblanc 276-278 

Non-contact — 

Surface 278-281 

Two-pass or double flow 280-281 

Ultimate analysis of coal 320, 322 

Underfeed stokers 348-354 

Uniflow engine 208-212 

Unit of heat energy 13 

Units of matter, energy and work 3 

Updraft furnace 315 

Utilization of exhaust steam for heating buildings 400, 401 

Vacuum 262-265 

Measurement of 262-265 

Pump 282-288 

Valve, D-slide — (see D-slide valve) 159-195 

Setting. 187-195 



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